Seal Systems, Bearing Arrangements and Couplings in Saudi A

Note: The source of the technical material in this volume is the Professional Engineering Development Program (PEDP) of Engineering Services.

Warning: The material contained in this document was developed for Saudi Aramco and is intended for the exclusive use of Saudi Aramco’s employees. Any material contained in this document which is not already in the public domain may not be copied, reproduced, sold, given, or disclosed to third parties, or otherwise used in whole, or in part, without the written permission of the Vice President, Engineering Services, Saudi Aramco.

Chapter : Mechanical For additional information on this subject, contact File Reference: MEX-211.04 PEDD Coordinator on 874-6556

Engineering Encyclopedia Saudi Aramco DeskTop Standards

SEAL SYSTEMS, BEARING ARRANGEMENTS, AND COUPLINGS IN SAUDI ARAMCO PUMPS

Engineering Encyclopedia Pumps

Seal Systems, Bearing Arrangements, and Couplings in Saudi Aramco Pumps

Saudi Aramco DeskTop Standards i

Section Page

INFORMATION ............................................................................................................... 5 INTRODUCTION............................................................................................................. 5 PUMP SEAL SYSTEMS.................................................................................................. 6

Packing................................................................................................................. 7

Compression Packing ................................................................................ 7

Molded Packing ....................................................................................... 15

Floating Packing ...................................................................................... 20

Inspection and Maintenance .................................................................... 25 Mechanical Seals ............................................................................................... 27

Basic Mechanical Seal Design................................................................. 29

Face Contact Seals.................................................................................. 33

Lift-Off Face Seals ................................................................................... 68

Auxiliary Seals ......................................................................................... 71 PUMP BEARING ARRANGEMENTS AND LUBRICATION REQUIREMENTS ............ 73

Bearing Loading ................................................................................................. 73

Radial Loads............................................................................................ 73

Axial Loads .............................................................................................. 76

Thrust Direction ....................................................................................... 80

Thrust Balancing Designs ........................................................................ 82 Bearing Types .................................................................................................... 92

Antifriction ................................................................................................ 93

Hydrodynamic.......................................................................................... 98

Pump Industry Standard Bearing Applications....................................... 108 Lubrication Requirements................................................................................. 114

Grease Lubricated ................................................................................. 114

Oil Lubricated......................................................................................... 115 PUMP COUPLINGS.................................................................................................... 118

Rigid Adjustable Spacer Type .......................................................................... 118 Flexible Disk Pack ............................................................................................ 120



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Diaphragm........................................................................................................ 121 Elastomeric....................................................................................................... 124

GLOSSARY ................................................................................................................ 127

LIST OF FIGURES

Figure 1. Typical Compression Packing Arrangement ................................................... 9 Figure 2. Styles of Braided or Twisted Packing............................................................ 12 Figure 3. Braided Packing Under Pressure .................................................................. 13 Figure 4. Metal Packing................................................................................................ 15 Figure 5. V-Ring Packing ............................................................................................. 16 Figure 6. U-Ring Packing Used to Seal a Piston.......................................................... 17 Figure 7. Cup Packing.................................................................................................. 18 Figure 8. Flange Packing ............................................................................................. 19 Figure 9. Expanding Split Piston Ring Seal .................................................................. 21 Figure 10. Contracting Split Ring Seal ......................................................................... 23 Figure 11. Different Types of Split Rings...................................................................... 24 Figure 12. Basic Mechanical Seal ................................................................................ 29 Figure 13. Basic Mechanical Seal-Improved ................................................................ 30 Figure 14. Basic Mechanical Seal with Static Seals ..................................................... 31 Figure 15. Basic Mechanical Seal with Spring ............................................................. 32 Figure 16. Mechanical Seal Construction (Type A, Pusher)......................................... 34 Figure 17. Mechanical Seal Sealing Points .................................................................. 35 Figure 18. Mating Ring Mounting Configurations ......................................................... 36 Figure 19. Single-Seal Arrangements .......................................................................... 38 Figure 20. Back-to-Back Double Seal Arrangement..................................................... 39 Figure 21. Opposed Dual Seal Arrangement ............................................................... 40 Figure 22. Tandem Seal Arrangement ......................................................................... 41 Figure 23. Single Spring Mechanical Seal.................................................................... 43 Figure 24. Multiple Spring Mechanical Seal ................................................................. 44 Figure 25. Bellows Mechanical Seal............................................................................. 45



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Figure 26. Wedge-Type Pusher Secondary Seal ......................................................... 46 Figure 27. O-Ring-Type Pusher Secondary Seal ......................................................... 46 Figure 28. Metal Bellows Nonpusher-Type Secondary Seal ........................................ 48 Figure 29. Elastomer Bellows Nonpusher-Type Secondary Seal ................................. 48 Figure 30. Unbalanced Seal......................................................................................... 49 Figure 31. Balanced Seal ............................................................................................. 50 Figure 32. Balancing Force on a Mechanical Seal ....................................................... 52 Figure 33. Balance Ratio Measurement Points ............................................................ 53 Figure 34. Flush Systems Using API 682 Plans........................................................... 62 Figure 35. Vendor Representations of Arrangement 1 Mechanical Seals .................... 64 Figure 36. Vendor Representation of a Type A, Arrangement 2 Mechanical

Seal ............................................................................................................ 65 Figure 37. Vendor Representation of a Type A, Arrangement 3 Mechanical

Seal ............................................................................................................ 66 Figure 38. Single, Dry-Gas, Lift-Off Seal ...................................................................... 68 Figure 39. Lift-Off Seal Rotating Face V-Groove.......................................................... 70 Figure 40. Wet Running Auxiliary Seal......................................................................... 71 Figure 41. Dry Running Auxiliary Seal.......................................................................... 72 Figure 42. Radial Force................................................................................................ 74 Figure 43. Radial Force Factor Coefficient Plot............................................................ 75 Figure 44. Hydraulic Axial Thrust Produced by a Horizontal, Single-Stage,

Single-Suction, Closed-Impeller Pump....................................................... 76 Figure 45. Additional Axial Thrust on an Overhung Pump............................................ 77 Figure 46. Hydraulic Axial Thrust Produced by a Horizontal, Single-Stage,

Double-Suction, Closed-Impeller Pump...................................................... 78 Figure 47. Axial Thrust in Horizontal, Single-Suction, Semi-Open, Radial

Flow Impellers ............................................................................................ 80 Figure 48. Examples of Directions of Axial Thrust........................................................ 81 Figure 49. Actual Pressure Distribution across an Impeller.......................................... 83 Figure 50. Front and Back Wear Rings and Balance Holes ......................................... 84 Figure 51. Pressure Differential across an Impeller with Pump-Out Vanes.................. 85 Figure 52. Stacked Impeller Design with Hydraulic Balancing Device.......................... 86 Figure 53. Opposed Impeller Design............................................................................ 87 Figure 54. Balancing Drum........................................................................................... 88



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Figure 55. Balancing Disk ............................................................................................ 90 Figure 56. Typical Pump Bearing Applications............................................................. 92 Figure 57. Single-Row Ball Bearing.............................................................................. 94 Figure 58. Double-Row, Angular-Contact Ball Bearing ................................................ 95 Figure 59. Angular-Contact Bearing ............................................................................. 96 Figure 60. Spherical Barrel-Shaped Roller Bearing...................................................... 97 Figure 61. Shaft/Bearing Dynamics.............................................................................. 99 Figure 62. Typical Journal Bearing............................................................................. 101 Figure 63. Multi-Lobe Journal Bearing ....................................................................... 102 Figure 64. Typical Tilting-Pad Bearing ....................................................................... 104 Figure 65. Tapered Land Thrust Bearing ................................................................... 105 Figure 66. Tilting-Pad Thrust Bearing......................................................................... 106 Figure 67. Self-Equalizing Thrust Bearing.................................................................. 108 Figure 68. Typical API Bearing Configuration ............................................................ 109 Figure 69. Constant-Level Oiler ................................................................................. 115 Figure 70. Typical Forced Feed Lubrication System .................................................. 117 Figure 71. Rigid Adjustable Coupling ......................................................................... 119 Figure 72. Double Disc-Pack Coupling....................................................................... 121 Figure 73. Diaphragm Coupling.................................................................................. 123 Figure 74. Elastomeric Couplings............................................................................... 126

LIST OF TABLES

Table 1. dN Rating/Bearing Type/Lubrication Type Table.......................................... 113



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INFORMATION

INTRODUCTION There are a few major components of a centrifugal pump, including the pump casing, the shaft, and the impeller. Other components that have a direct effect on the successful installation and operation of a centrifugal pump include shaft seals, bearings, and couplings. Pump sealing systems, bearings, and couplings are commonly responsible for the majority of centrifugal pump downtime and repair. The specific design and material consideration for pump shaft seals, bearings, and couplings used in Saudi Aramco installations will vary based on the specific application, the type of fluid being pumped, and the type of pump used for the given application. The Mechanical Engineer must understand the mechanical aspects of pump shaft sealing systems, bearings, and couplings to make the proper selection of these components for a pump installation. In this section, the following mechanical aspects of pumps will be described:

• Pump Seal Systems

• Pump Bearing Arrangements

• Pump Couplings




PUMP SEAL SYSTEMS The use of proper seals at mechanical joints and along shafts that rotate is a primary concern for the Mechanical Engineer. A seal is a device or material that controls the leakage of fluids from a pump by creating and/or maintaining a fluid-pressure differential across the gap that exists between two relatively movable and/or separable components of a fluid system. Prevention of leakage is necessary for the following reasons:

• To prevent loss of the pumpage to the drain system (economical).

• To prevent leakage of the pumpage to the atmosphere (safety and environmental).

• To prevent damage to rotating equipment by the prevention of leakage from one section of a piece of rotating equipment (high pressure area) to another section of the piece of rotating equipment (low pressure area).

• To prevent contamination of clean sections of a piece of rotating equipment by the prevention of leakage of contaminated pumpage to the clean sections of the rotating equipment.

Seals involve an almost unlimited variety of sizes, configurations, materials, and material combinations. This section will examine the following forms of sealing devices for use on pumps:

• Packing

• Mechanical Seals

The need to control leakage should determine which type of shaft seal is selected. Where leakage is acceptable, such as in raw water, seawater, and potable water services, seals with a higher leakage rate are acceptable. In general, leakage from hydrocarbon and chemical services must be minimized, so face-type mechanical seals or sealless pumps are commonly used.




Packing Packing is a method used to seal the rotating or reciprocating shaft of a pump through the use of rings of flexible material that are placed around the shaft. Packing is a type of seal that is used to prevent or restrict leakage of a contained liquid around or along a shaft, a plunger, a piston, a ram, or through a mechanical joint. Packing also is used to stop contaminants from entering a system.

Packing that is installed in a rotary or reciprocating joint is classified as a dynamic seal. Packing that is installed on a stationary joint is classified as a static seal. In accordance with SAES-G-005, packing is required for main firewater service pumps, for pumps in oily water or storm water sump service, and for slurry service. Packing may also be used to seal the piston or plunger of a reciprocating power pump. Because the mechanical joints on pumps can be subjected to various system conditions (e.g., temperature, pressure, shaft speed, and process fluid), each type of packing is available in a wide variety of materials and configurations. The following three types of packing will be discussed in this section:

• Compression packing

• Molded packing

• Floating packing

Compression Packing

For many years, compression packing has been used to seal rotating equipment. For compression packing to provide a seal, the compression packing must be radially distorted so that it is in contact with the sealed surfaces (seal chamber and shaft). Such radial distortion is achieved through use of a packing gland follower that mechanically applies an axial pressure to the packing.




Figure 1 shows a typical compression packing arrangement for a centrifugal pump and an expanded view of the seal chamber region. Figure 1 shows the seals that are formed though compression of the packing material that is between the packing gland follower and the seal chamber throat. The packing gland follower creates compression forces when the packing gland stud is tightened. These compression forces axially push the compression packing material against the throat of the seal chamber and radially against the shaft and seal chamber. A static seal is formed at the inside diameter of the seal chamber, and a dynamic seal is formed between the compression packing material and the rotating shaft or shaft sleeve.

As shown in Figure 1, some packing arrangements use a lantern ring to allow the use of an external seal fluid in the seal chamber. A lantern ring, also called a seal cage, is installed between the packing at the sealing fluid connection. Water or other sealing fluid is introduced under pressure into the seal chamber through the lantern ring. To prevent blocking of the lantern ring from the seal fluid connection, care must be taken when tightening packing on pumps that use lantern rings.

Some leakage along the dynamic seal of the shaft is necessary to cool and lubricate the packing. The amount of leakage necessary for such cooling and lubrication depends on the quality of the material, the operating conditions of the application, and the condition of the equipment. When excessive leakage occurs along the shaft, the packing gland is further tightened.

Excessive leakage is any amount of leakage flow that exceeds the amount of flow necessary to cool and lubricate the packing. Excessive leakage should be avoided to prevent the loss of the sealed liquid to either a drain system or to atmosphere. Excessive leakage also should be avoided because it can compound itself in that excessive leakage will wash the lubricant out of the packing material and cause increased leakage. As the lubricant is washed out of the packing material, an amount of torque will be required to tighten the packing. The increased torque can cause the packing to overheat and burn.




Figure 1. Typical Compression Packing Arrangement




Because the packing will wear as rotating equipment is operated, the packing gland loading (compression forces) on the compression packing must be adjusted to control leakage. As the packing gland loading is increased, the packing dimension is squeezed smaller and smaller. Care should be taken not to overtighten the packing. Overtightening of compression packing material shortens the life of the packing material. Excessive overtightening will destroy the sealing capability of the packing, and it can cause the portion of the equipment that rotates to seize. The lubricant that is incorporated in compression packing prevents burning and scoring of the shaft. Some of the lubricant will be driven from the packing after repeated packing gland loading adjustments. This loss of lubricant will reduce the operational life of the packing.

Compression packing commonly used in pumps come in several construction configurations. The following types of configurations are discussed below:

• Braided or Twisted Packing

• Plastic Packing

• Metal Packing

Saudi Aramco Standard 31-SAMSS-004 prohibits the use of asbestos in pump packing materials.

Braided or Twisted Packing - Braided or twisted packing strands are individually impregnated with mineral oil, grease, or graphite to retain the packing flexibility, to lubricate parts that move, and to help create a fluid seal under pressure. Figure 2 shows the following four basic styles of braided or twisted packing:

• Twisted braid packing - In this style of packing, yarns are twisted around each other to obtain the desired size. Because the strands can be untwisted and removed, one packing size can be used in seal chambers of various sizes. Such flexibility allows the desired size to be custom-made through use of only the number of strands that are needed. Twisted packing is used in general utility or emergency-type packing for pumps in which packing space is small.




• Square braid packing - In this style of packing, each strand passes over and under strands that are running in the opposite direction. The cross-section of this packing is square. Square braid packing is preferred for high-speed rotary and reciprocating service.

• Braid-over-braid packing - In this style of packing, packing is built up to the required size by braiding one or more covers around a central core of braided or twisted homogeneous materials.

• Interlocking braid packing - In this style of packing, diagonally braided packing is braided inside as well as outside. Each strand diagonally passes through the body of the packing at an angle of approximately 45°. This pattern produces a completely unified structure. Each braided strand contributes to the strength of the entire packing. Because diagonal braiding makes each strand much more flexible than ordinary braiding, there is less stress when the packing is formed into rings. Because of the increased flexibility of this packing, this style of packing is ideal for pumps with small diameter, high speed shafts.




Figure 2. Styles of Braided or Twisted Packing




Figure 3 shows an exploded view of one ring of braided packing that is installed in a seal chamber. The packing gland load causes the braided packing to create a sealing force against the seal chamber and the shaft. This force effectively seals against system pressure.

Figure 3. Braided Packing Under Pressure




Plastic Packing - Plastic packing are used in several forms. These forms physically appear the same as those packing that were previously shown in Figure 2. One form of plastic packing consists of compounded material that is combined with binding materials and lubricants. The combination of materials then is extruded into a rectangular cross-section in spiral form. The packing that is produced is soft and readily formable, and it is particularly suitable for pumps.

Two other common forms of plastic packing are the various Teflon (TFE) packing and TFE-impregnated packing. TFE packing are produced from a continuous TFE filament in the same style as the conventional braided and twisted packing. After braiding, some styles are impregnated with a TFE suspension to produce a 100 percent TFE packing. TFE-impregnated packing are formed braided packing that are impregnated with TFE. For extremely critical services, each strand is impregnated before the strand is braided. After the strand is braided, each braid is impregnated again with TFE suspension. Caution should be exercised in the application of certain types of TFE packing. Teflon has a temperature limit of 500ºF. Teflon can break down and creep.




Metal Packing - Metal packing must be flexible, compressible, and have little friction. Metals are crimped, spiral wrapped, folded, and braided to form packing of a desired shape. Figure 4 shows a spiral-wrapped metal-foil packing and a folded and twisted metal-foil packing.

Figure 4. Metal Packing

Molded Packing

Molded packing sometimes are referred to as automatic, hydraulic, or mechanical packing. Molded packing rely on fluid pressure to force the packing material against the wear surfaces. These packing are made from leather, rubber, fabric-reinforced rubber, or synthetic rubber. The following are the three general types of molded packing:

• V-ring packing

• U-ring packing

• Lip-type packing




V-ring and U-ring packing can be used in pumps that move fluids at both high pressures (up to 50,000 psi) and low pressures. Lip-type packing primarily are used in pumps with reciprocating shafts

V-Ring Packing - Multiple V-rings are often installed in sets, and they are packed on the outside of a reciprocating shaft, as shown in Figure 5. A support ring or adapter is used in the bottom of the seal chamber to support the V-ring packing. The shape of the support ring and packing gland follower maintains the V-ring packing shape. The seal is created by the close fit of the packing.

Figure 5. V-Ring Packing




U-Ring Packing - U-ring packing usually is used to seal a piston, as shown in Figure 6. U-ring packing uses the system pressure to assist in the seal. When the piston is on its compression stroke, the system pressure will be felt in the U-ring groove of the piston, and thereby cause the U-ring to seal against the cylinder wall.

Figure 6. U-Ring Packing Used to Seal a Piston




Lip-Type Packing - The extended lip section is what differentiates lip-type packing from V-ring and U-ring packing. Lip-type packing are further classified as either cup or flanged. Cup packing have a single lip, and they are used to seal pistons, as shown in Figure 7. Cup packing is held in place and directly attached to the base of a reciprocating piston through use of a packing retainer plate. The cup packing uses system pressure to assist in the seal in the same manner as the U-ring packing.

Figure 7. Cup Packing




Flange packing - Flange packing, as shown in Figure 8, is another version of lip-type packing. Flange packing is used to seal shafts in low-pressure applications in which sufficient room does not exist for V- or U-ring style packing. Flange packing is preformed to match a specific piece of equipment or a specific set of dimensions. Flange packing is usually molded from tough materials, such as carboxylated nitriles and polyurethanes that resist abrasion.

The sealing action of the lip is actuated by the process pressure that forces the lip against the shaft. Clearance must exist between the lip and the bottom of the seal chamber. If the lip strikes the bottom of the seal chamber, the lip will turn outward and destroy the flange packing ability to seal. The flange section of the packing is sealed by a compression force that is supplied by the gland. The compression force must be strong enough to seal against the maximum process pressure that will be applied.

Figure 8. Flange Packing




Floating Packing

Floating packing often is called split ring seals. Floating packing is designed to move as the component that it is sealing moves. There are two basic types of floating packing: expanding and contracting. Expanding split piston ring seals expand outward toward the cylinder wall to provide the seal for the fluid in pumps that move liquids at high pressures or high temperatures. Contracting split ring or rod seals contract inward toward the reciprocating rod, and they are used whenever space, high temperature, or excessive pressure prohibit the use of other packing.

An example of an expanding split piston ring seal is shown in Figure 9. In all applications, the process pressure is felt on the split ring and forces the ring against the surfaces that require sealing. Expanding split ring seals must mate on the inner diameter of the fixed cylinder wall (the primary contact point) and the top or the bottom (the secondary contact point) of the packing space.

As the reciprocating piston moves upward, the expanding split ring seal sits on the bottom of the packing space and seals the piston's edge. Process pressure pushes the expanding split piston ring seal out toward the fixed cylinder wall, and it increases the seal pressure at the primary contact point. The process pressure also pushes the expanding split ring seal down toward the bottom packing space surface, and it increases the seal pressure at the secondary contact point.




As the reciprocating piston moves downward, the expanding split ring seal sits on the top of the packing space and still seals the piston. The packing, however, has "floated" within the packing space region to a new position. The primary contact point remains the same, which is the inner diameter of the fixed cylinder wall; however, the secondary contact point has moved from the bottom of the packing space region to the top of the packing space region.

Figure 9. Expanding Split Piston Ring Seal




An example of a contracting split ring seal is shown in Figure 10. The contracting split ring seal performs its function in the same manner as the expanding split ring seal. Process pressure is felt on the split ring and forces the ring against the surfaces that require sealing. Contracting split ring seals must mate on the outer diameter of the reciprocating rod (the primary contact point) and the top or the bottom (the secondary contact point) of the packing space in the seal chamber.

As the reciprocating rod moves upward, the contracting split ring seal seals at the bottom of the packing space and seals the reciprocating rod's edge. Process pressure pushes the contracting split ring seal in toward the reciprocating rod, and it increases the seal pressure at the primary contact point. The process pressure also pushes the split ring down toward the bottom packing space surface, and it increases the seal pressure at the secondary contact point.

As the rod moves downward, the contracting split ring seal sits on the top of the packing space and still seals the rod; however, the packing has "floated" within the packing space region to a new position. The primary contact point remains the same, which is the inner diameter of the fixed cylinder wall; however, the secondary contact point has moved from the bottom of the packing space to the top of the packing space.




Figure 10. Contracting Split Ring Seal

Figure 11 shows the following types of split rings:

• The simple straight-cut split ring that is used as a low-pressure piston seal for applications in which joint leakage is not critical.

• The step-cut split ring that is used to reduce joint leakage.

• The balanced ring that is used when hydraulic fluid pressure is high.




Figure 11. Different Types of Split Rings




Inspection and Maintenance

The packing should be able to perform its function without rapid deterioration of the packing or wear on rotating components; however, there is still some amount of friction that is created between the shaft and the packing. This friction is the major cause of packing and shaft sleeve wear. The packing eventually will need to be replaced when leakage from the gland exceeds the acceptable levels. The major causes of rapid packing wear due to excessive friction include the following:

• Shaft sleeve runout

• Misalignment of the shaft with the seal chamber

• Lack of packing lubricant and cooling

• Abrasives that are trapped between the shaft and the packing

• Overtightening the packing

Shaft sleeve runout will cause localized heating due to increased friction on the runout portions of the shaft. The higher friction areas will wear the packing significantly faster than other portions of the packing. These areas will lead to excessive packing leakage.

Misalignment of a shaft with the seal chamber will cause uneven wear on a localized portion of the packing. Because the packing is stationary, the same area of packing will bear an increased load that is caused by the misalignment. The increased load generates more friction and results in a more rapid breakdown of the packing.

Abrasives that are trapped between the shaft and the packing result in a scratched or worn packing and shaft sleeve. Packing and sleeve damage is common. The scratches will initially be filled by lubricating mediums or the packing itself due to expansion; however, too many abrasives will rapidly wear away too much packing to the point that the packing can no longer seal the shaft.




The use of packing is maintenance intensive and regular inspections must be performed. The life of packing can be extended through use of inspection and packing adjustments; however, if there is no regular inspection of the packing to detect excessive leakage, the leakage will continue to increase and cause a faster deterioration of the packing.

Packing adjustments must be performed in accordance with appropriate procedures. Each style of packing has its own set of requirements that require specific methods of adjustment. Failure to follow the appropriate procedure can lead to more damage to the packing and to increased leakage rather than reduced leakage.

The inspections should include leakage rates, fraying, crushed or pinched packing, and glazing. Some leakage is required to lubricate and to cool the rotating shaft. Minor fraying indicates that the packing compression is high and that the packing will need to be replaced soon. Crushed or pinched packing indicates misalignment of the rotating shaft or excessive gland nut pressure. Both of these conditions lead to packing replacement. Glazed packing indicates that the packing is too tight and that the friction between the rotating portion and the packing is overheating and is, as a result, destroying the packing.

Proper documentation of the amount of packing leakage in the maintenance records will allow the Mechanical Engineer to determine trends in packing performance. These trends can be used to predict how often the packing should be inspected and replaced. The records can also indicate a need to alter the packing style or the methods for which packing is used for the application.




Mechanical Seals Similar to packing, pump mechanical seals minimize leakage of liquids or gases in rotating shaft applications. A mechanical seal is an end-face seal that is designed to provide rotary seal faces that can operate with no visible leakage. All mechanical seals leak. Mechanical seals operate on the principle that the seal design vaporizes the liquid in the seal chamber across the primary seal faces (approximately 50 to 75 percent down the seal face width). The vaporization of liquid across the seal faces results in a small, undetectable leakage of the fluid (in its vapor state) from the seal chamber. Almost all mechanical seal designs use two mating rings, one ring rotates and the other ring is stationary, to provide sealing surfaces at the point of relative movement. The mechanical seal design includes a closing force to hold the seal faces in contact. This force is supplied by a combination of system pressure and spring force, each of which is discussed in detail later in this section.

Mechanical seals are installed in the majority of the process pumps at Saudi Aramco. Mechanical seals may be used on pumps when the nature of the liquid handled by the pump is hazardous.

Sealless pumps are required for low horsepower chemical, sour water, and acid service.

Mechanical seals will give better service than conventional packing. Mechanical seals may be used in a conventional pump seal chamber to seal any number of liquids at various speeds, pressures, and temperatures. Saudi Aramco requires mechanical seals for all hydrocarbon and special-purpose water services, such as boiler feed pumps and water injection. The following comparison between mechanical seals and packing illustrates why mechanical seals are used in Saudi Aramco process systems:

• Pollution - Environmental agencies will no longer tolerate process pump leakage polluting water sources and soil. Packing leakage is greater than mechanical seal leakage: therefore, more leakage waste occurs with packing. Treatment of this waste is very expensive.




• Cost of product - A steady drip from packing will equal 20 to 25 gallons per day. The cost of many products will more than pay for the seals needed to stop the leakage.

• Power consumption - A mechanical seal consumes approximately one-sixth the electric power of packing. Packing rubbing on a shaft or sleeve is similar to driving an automobile with the parking brake engaged.

• Cost of bearings - Most pump bearing failure is caused by contamination rather than overloading. The easiest way to contaminate a bearing is from the leakage coming through the packing. Reduction of leakage by using a mechanical seal can reduce bearing failures.

• Cost of sleeves or shafts - Packing will damage shafts or shaft sleeves because of the direct contact between the rotating shaft and the stationary packing. There is no contact between the rotating shaft and the nonrotating mechanical seal parts.

• Packing advantages - Packing can be more advantageous in conditions where product leakage or product cost is not an issue, such as fire water pumps or sewage pumps. Packing is also less expensive than mechanical seals, and it is less prone to catastrophic failure.

In accordance with 31-SAMSS-004, mechanical seals on pumps with shaft sizes from 1.5 to 4.5 inches (30 to 120 mm) must be installed in accordance with the following standards:

• API Standard 682 - Shaft Sealing Systems for Centrifugal and Rotary Pumps

• Saudi Aramco Standard 31-SAMSS-012 - Shaft Sealing Systems for Centrifugal and Rotary Pumps

Saudi Aramco Standard SAES-G-005 provides a mechanical seal selection guide (Table 1 of SAES-G-005) that aids in identifying what type of seal is required for the various applications at Saudi Aramco facilities.




Basic Mechanical Seal Design

The simplest form of a mechanical seal, shown in Figure 12, has a shoulder on the pump shaft that presses against the machine housing. This configuration will only work if the housing and the shoulder are both finished properly and the shaft is loaded against the housing equally at all times. This configuration is not practical on most machines because there is always some shaft end play and seal face run-out. This configuration also does not allow for the adjustment of the impeller endplay on open-faced impeller pumps. Wear can reduce the clearance between the shaft and the housing, which will result in increased leakage.

Figure 12. Basic Mechanical Seal




If a sealing ring is installed on the shaft and mating ring on the housing, as shown in Figure 13, the wear will take place on the sealing rings and not on the shaft and housing. The sealing ring surfaces are easier to machine (lap) to maintain uniform flatness of the mating surfaces. Also, the rings can be renewed when worn and without machining the shaft or housing. However, keeping the shaft against the sealing area is still a problem. Shaft run-out and endplay are still a problem because the rings are rigid and seal face run-out would cause excessive leakage.

Figure 13. Basic Mechanical Seal-Improved




Mounting a static seal in back of the sealing and mating rings, as shown in Figure 14, will allow some movement of sealing rings, which also seals fluid leakage around the back of the sealing rings. Shaft end motion is still limited with this design because there is only so much motion the gasket material can absorb. This design still will not maintain correct contact force to keep the seal faces together.

Figure 14. Basic Mechanical Seal with Static Seals




If a spring is added to the seal behind either the shaft sealing ring or the mating ring, as shown in Figure 15, contact between the sealing rings can be maintained. The spring will absorb any movement of the shaft from endplay, thermal expansion, and run-out, and it will automatically adjust the seal for wear. The mating ring is installed in a gland that facilitates seal removal from the pump during maintenance

Figure 15. Basic Mechanical Seal with Spring

The design of a mechanical seal can be a face contact seal type or lift-off face seal type. The following sections describe the types of mechanical seals, construction, classifications, and flushing/cooling systems.




Face Contact Seals

An example of a face contact seal is shown in Figure 16. Mechanical seal sealing surfaces are perpendicular to the shaft, with contact between the seal ring and the mating rings to achieve a dynamic seal. The seal ring can be flexibly mounted in the seal head assembly, which usually rotates with the shaft. The mating ring is usually fixed to the pump gland plate and does not rotate with the shaft. Each of the sealing planes is lapped flat within two to three light bands. Wear occurs at the dynamic sealing faces from sliding contact between the seal and mating rings. The amount of wear is small, as a film of the liquid sealed is maintained between the sealing faces during pump operation. The mating surfaces of the sealing faces are held in contact by a spring and the fluid forces present in the seal chamber. The preload of the spring is required to produce the initial seal during shutdown or when there is a lack of fluid pressure behind seal. During operation, the spring and fluid pressure behind the seal maintains the contact between the sealing faces.

All mechanical seals have two assemblies. The seal head assembly is made up of the seal ring, a spring or metal bellows (not shown), and a secondary seal between the seal ring and the shaft (type A seals only). The mating ring assembly is made up of the mating ring and a static seal between the mating ring and the pump gland plate.




Figure 16. Mechanical Seal Construction (Type A, Pusher)

As shown in Figure 17, the following three points of sealing are common to all mechanical seal installations:

• Seal Point 1: At the mating surfaces of the primary and mating rings.

• Seal Point 2: Between the seal head assembly and the shaft or the shaft sleeve.

• Seal Point 3: Between the mating ring assembly and the pump gland plate.




Figure 17. Mechanical Seal Sealing Points

The seal between the seal head assembly and the shaft or shaft sleeve is called the secondary seal. The secondary seal must be dynamic to maintain the seal as the sealing ring faces wear, as the shaft moves from runout, as the shaft and pump casing expands from thermal expansion, and as the machinery vibrates. Secondary seals are commonly an O-ring or graphite. The selection of secondary seals will depend on the type of service (described later in this module).

The mating ring/gland plate seal is a static seal. The mating ring/gland plate seal is typically an O-ring or a graphite gasket.

The mating ring must remain stationary in the gland plate. The following are three common methods to hold the mating ring stationary:

• The pressure loading of the mating ring seal (O-ring or cup) against the gland plate.

• Pinning the gland plate to a recess in the mating ring.

• Utilizing a clamped-in design, which is prohibited by 31-SAMSS-012.

Figure 18 shows the different methods used to hold the mating ring stationary.




Figure 18. Mating Ring Mounting Configurations




Face seals can be classified by the following:

• Arrangement

• Secondary seal design

• Primary ring balance

Single-Seal Arrangement - As shown in Figure 19, the single seal is the most common mechanical seal and is used in non-hydrocarbon and many low vapor pressure hydrocarbon applications. A single seal is the simplest seal arrangement. The single seal may be internally mounted inside the seal chamber, or it may be externally mounted outside the seal chamber. Internally mounted seals use liquid under pressure in the seal chamber acting with the spring load to keep the seal faces in contact. Externally mounted seals are not used in Saudi Aramco applications.




Figure 19. Single-Seal Arrangements




Dual Seal Arrangement - Dual seals are used in applications that require a neutral liquid for lubrication, improved corrosion resistance, or an increase in plant safety. Dual mechanical seals may be installed in a back-to-back arrangement, a face-to-face arrangement, or a tandem arrangement. Figure 20 shows a back-to-back arrangement. The back-to-back arrangement consists of two single seals back-to-back with the primary rings facing in opposite directions in the seal chamber. Each primary ring has its own mating ring. The seal may have a single spring or multiple springs located between the primary rings. Barrier fluid is applied to the gland seal area between the primary rings at a higher pressure than that of the liquid being pumped. This barrier lubricates the sealing faces and applies hydraulic pressure to the primary rings. The inboard seal keeps the liquid being pumped from entering the seal chamber. This type of arrangement is used when there is the possibility of the pumped fluid entering the atmosphere.

Figure 20. Back-to-Back Double Seal Arrangement (Only used in Brine Water Injection Pumps)




Figure 21 shows an example of a face-to-face arrangement. The opposed arrangement consists of two single seals mounted face-to-face. The two seal rings are seated against a single mating ring. Each primary ring has its own spring or metal bellows (not shown). Buffer fluid is applied between the seals. The inboard seal carries the full differential pressure of the pump seal chamber to the buffer fluid. The outboard seal carries only the differential pressure of the buffer fluid pressure to the atmosphere. This arrangement allows for seal installation with a shorter axial length than is possible with the back-to-back arrangement while still increasing the reliability for plant safety.

Figure 21. Opposed Dual Seal Arrangement (Dual Pressurized Arrangement 3)




Figure 22 shows an example of a tandem (dual pressurized arrangement 3) mechanical seal arrangement. Tandem mechanical seals consist of two single mechanical seals facing in the same direction. There are two separate seal rings, mating rings (in each gland plate), and springs or metal bellows (not shown) located in the seal head. Buffer fluid is applied to the outboard seal and creates a buffer zone between the liquid being pumped and the atmosphere. Normally, the pressure differential between the liquid being pumped and the atmosphere is taken across the inboard seal, with the buffer fluid at or near atmospheric pressure. The liquid in the outboard seal chamber may be circulated to remove seal heat. The inboard seal may be lubricated and cooled from the pump discharge or an external flush source if the pumped fluid is not clean.

Figure 22. Tandem Seal Arrangement




Rotating or Stationary Seal Head - The rotating seal head design (API Arrangement 1) is the most common seal type. The seal head assembly rotates with the pump shaft. The rotating seal head design is limited by pump shaft speed and is usually not used for shaft speeds greater than 5000 revolutions per minute.

The stationary seal head (API 682 Type C Seal) has the mating ring assembly rotating with the pump shaft. The advantage to this design is that the stationary flexible element does not experience any distortion due to rotational forces. The mating ring assembly is more capable of resisting rotational forces because of its compact, rigid design. The mating ring assembly is easier to dynamically balance and therefore is used on pumps that operate at high speeds (greater than 5000 revolutions per minute).




Single or Multiple Spring Construction - The single spring construction seal is not the preferred design for petrochemical services. Figure 23 shows an example of a single spring mechanical seal design. This design consists of a single large-diameter spring. This design has the advantages of withstanding a large amount of corrosion, and the openness of the design prevents clogging. The disadvantage of this design is that the seal requires a deep seal chamber for the installation, and the single spring may not give even pressure on the primary ring at high speeds, and seal face tracking is poor.

Figure 23. Single Spring Mechanical Seal




Figure 24 shows the multiple spring construction seal. Small individual springs are placed around the circumference of the primary seal. The advantages of this design are that the seal can be installed in pumps with short seal chamber depth and the spring pressure on primary ring is evenly distributed. The disadvantages of this design are that the small spring wire is more susceptible to failure from corrosion and the small springs clog, which defeats the function of the spring. However, almost all petrochemical sealing services are better served with a multiple spring seal design. The multiple spring design is a requirement of API 682 and 31-SAMSS-012.

Figure 24. Multiple Spring Mechanical Seal

Some seal designs utilize a metal bellows instead of a spring. Figure 25 shows an example of a bellows mechanical seal. The metal bellows is equivalent to a single large spring and has several advantages over the spring design. The metal bellows is able to drive the primary ring without the use of a dynamic secondary seal. The disadvantage of this design is that the bellows has pressure limitations, usually limited to pressures below 250 psig. This design is also expensive and is sensitive to twisting and rupture if the seal faces stick.




Figure 25. Bellows Mechanical Seal

Saudi Aramco Standard 31-SAMSS-012 requires that mechanical seal springs, single or multiple, be constructed from Hastelloy C-276, Inconel 625, or Inconel 718.




Pusher and Nonpusher Seals - A pusher secondary seal design is defined as seal assemblies in which the secondary seal is moved along the shaft by mechanical load and hydraulic pressure in the seal chamber. Two examples of pusher seal secondary seals are shown in Figures 26 and 27. Pusher seal secondary seals are typically an O-ring, a wedge type, or a V-ring design. Figure 26 is a wedge-type pusher secondary seal. Figure 27 shows a conventional O-ring-type pusher secondary seal. The mating ring is stationary while the primary ring rotates with the shaft. The seal chamber pressure and spring or springs load the primary ring.

Figure 26. Wedge-Type Pusher Secondary Seal

Figure 27. O-Ring-Type Pusher Secondary Seal




Nonpusher secondary seal design is defined as seal assemblies in which the secondary seal is stationary, not dynamic (forced along the shaft by mechanical load). Seal movement is taken up by convolution of a bellows. Figure 28 shows a metal bellows nonpusher-type secondary seal. Figure 29 shows an elastomer bellows nonpusher-type secondary seal.

It is important to note that since the bellows provides the flexibility of the spring but does not leak, the secondary seal is static and not dynamic. Dynamic secondary seals can hang up if dirt accumulates adjacent to the O-ring.

Saudi Aramco Standard SAES-G-005 states that a bellows seal can be an acceptable alternative to a standard pusher seal in applications where the maximum seal chamber pressure does not exceed 1830 kPa (265 psia).

Justification for the selection of a bellows seal instead of a pusher seal for low temperature services may include:

• standardization with existing seal installations

• good operating experience with similar seals

• elimination of the dynamic O-ring secondary seal

• elimination of the stepped shaft sleeve

• elimination of the requirement for an overlaid shaft sleeve under the dynamic O-ring.




Figure 28. Metal Bellows Nonpusher-Type Secondary Seal

Figure 29. Elastomer Bellows Nonpusher-Type Secondary Seal




Balanced and Nonbalanced Seals - The effectiveness of a mechanical seal is determined by the performance of the contact between the mating seal surfaces. If the seal load is too high, the liquid film between the seal rings could be squeezed out or vaporized, which results in a high rate of seal face wear. Balance is a way of controlling the contact pressure between the seal faces and the power losses by the seal. A seal may be either an unbalanced or a balanced type.

An unbalanced seal is shown in Figure 30. With an unbalanced seal, pump seal chamber pressure acts on the total area of the primary ring. The force is in one direction against the mating ring, which results in a high closing force. Unbalanced seals are only suitable for low-pressure applications generally less than 100 psig, and they are not acceptable under API 682.

Figure 30. Unbalanced Seal




A balanced mechanical seal is shown in Figure 31. With a balanced, pusher-type, mechanical seal, the shaft or shaft sleeve is shouldered and the primary ring sealing face is stepped. This arrangement provides an area for the seal chamber pressure to act on the front and back of the primary seal. The amount of closing force from the seal chamber pressure on the primary seal ring varies by varying the amount of the step on the primary ring. The force can be adjusted as required to control the friction on the seal faces for various liquids and shaft speeds. A balanced seal generates less heat, requires less power to operate, and will last longer than an unbalanced seal. A balanced seal is typically used for high-pressure applications.

Figure 31. Balanced Seal




The amount of seal balance is part of the basic seal design called the balance ratio. The seal balance ratio is the comparison between the seal area exposed to the force pressing the seal faces together and the seal area exposed to the force pushing the seal faces apart. Figure 32 illustrates balance ratio. The pressure, p, in the seal chamber acts equally in all directions and forces the primary ring against the mating ring. Pressure acts only on the annular area, ac, so that the closing force Fc on the seal face is as follows:

F+pa=F scc

Where:

Fc = Closing force in pounds force or Newtons

p = Seal chamber pressure in (psi, N/m2)

ac = Hydraulic closing area (in2, m2)

Fs = Spring or bellows force

The pressure between the primary and mating rings is determined by the ratio of the closing force, Fc, and the hydraulic opening area, ao, as follows:

apa

aFP

oc

oc ==

Where:

P = The pressure between the primary ring and mating ring

Fc = Closing force in pounds force or Newtons

p = Seal chamber pressure (psi, N/m2)

ac = Hydraulic closing area (in2, m2)

ao = Hydraulic opening area (in2, m2)




LIQUID

SEAL GLAND

ATMOS.

p

Fc ac

ao

LIQUID

SEAL GLAND

ATMOS.

p

Fc

ac ao

EQUAL AND OPPOSITE FORCES

UNBALANCED SEAL

BALANCED SEAL

Figure 32. Balancing Force on a Mechanical Seal

The amount of force present at the seal face can be varied by changing the magnitude of the opening and closing forces. If the closing area (ac) is decreased by adding a shoulder on a shaft sleeve or seal hardware, the seal face pressure can be lowered. A balanced pusher seal is designed to operate with the shoulder. Bellow seals are balanced by design.




Figure 33 shows the seal balance ratio parameters for a seal pressurized at the outside diameter and for a seal pressurized at the inside diameter. The seal balance ratio can be calculated as follows:

For seals pressurized at the outside diameter:

IDODBDODRatio Balance Seal

22

22

−−

=

For seals pressurized at the inside diameter:

IDODIDBDRatio Balance Seal

22

22

−−

=

Where:

OD = Seal face outside diameter

ID = Seal face inside diameter

BD = Balance diameter of the seal

DO = Dynamic O’Ring

OUTER DIAMETERPRESSURE

BD ODID

DO

INNER DIAMETERPRESSURE

BDOD

ID

DO

Figure 33. Balance Ratio Measurement Points

Balance diameter can vary with seal design. For spring pusher seals that are under outer diameter pressure, the balance diameter is the diameter of the sliding contact surface of the inner diameter of the dynamic O-ring. For spring pusher seals that are under inner diameter pressure, the balance diameter is the diameter of the sliding contact surface of the dynamic O-ring outer diameter.




Flushing/Cooling Requirements - The dynamic operation of a mechanical seal results in heat generation. The amount of heat generated by a mechanical seal for a given application is critical to the proper selection of the seal design, the materials, and the auxiliary system. The wear of mechanical seal faces is a direct function of the seal face temperature. The heat generated at the seal face can also cause thermal distortion of the seal face materials, which will result in seal leakage or failure. The amount of heat generated by a mechanical seal has a direct impact on the type of cooling system and seal face materials selected for a mechanical seal application. API Standard 682, Appendix H, provides guidance for calculating the heat generated at the seal faces.

The following sequence of equations is used to determine the heat generated by a seal:

1. The seal face area is determined as follows:

( )IDOD4A 22 −×

=π

Where: A = Seal face area OD = Seal face outer diameter in inches ID = Seal face inner diameter in inches

2. The seal balance ratio is determined as follows:

For seals pressurized at the outside diameter:

IDODBDODB

22

22

−−

=

For seals pressurized at the inside diameter:

IDODIDBDB

22

22

−−

=

Where: B = Seal balance ratio OD = Seal face outside diameter ID = Seal face inside diameter BD = Balance diameter of the seal




3. The seal spring force is determined as follows:

AF

P spsp =

Where:

Psp = Seal spring pressure in pound per square inch

Fsp = Spring force at working length (for multiple spring design, the value for one spring is used)

A = Seal face area

4. The total face pressure in pounds per square inch is determined as follows:

( ) PKBDP spptot +−×=

Where:

Ptot = Total face pressure in psi

Dp = Pressure across the seal face in psi (differential fluid pressure from the seal chamber to the opposite side of the mating ring)

B = Seal balance ratio

K = Pressure drop coefficient, assumed to be 0.5

Psp = Spring pressure

5. The mean face diameter in inches is calculated as follows:

( )2

IDODMD +=

Where:

MD = Mean face diameter in inches

OD = Seal face outer diameter in inches

ID = Seal face inner diameter in inches




6. The seal running torque is calculated as follows:

2MDfAPRT tot ×××=

Where:

RT = Seal running torque in inch-pounds

Ptot = Total face pressure in psi

A = Face area in square inches

f = Coefficient of friction

MD = Mean face diameter in inches

The value of the coefficient of friction (f) for various seal face materials is as follows:

Sliding Materials

Rotating Stationary Coefficient of friction (f)

Carbon-graphite (resin filled)

Cast iron Ceramic Tungsten carbide Silicon carbide Silicon carbide converted carbon

0.07 0.07 0.07 0.02 0.015

Silicon carbide Tungsten carbide 0.02

Silicon carbide converted carbon 0.05

Silicon carbide 0.02

Tungsten carbide 0.08




7. The seal power (hp) is calculated as follows:

( )63,000

NRThp ×=

Where:

hp = Seal power in horsepower

RT = Seal running torque in inch-pounds

N = Seal rotational speed in rpm

8. The heat generated by the seal is calculated as follows:

2545hpQheat ×=

Where:

Qheat = Heat generated by the seal in BTUs per hour

hp = Seal power in horsepower

The following example illustrates the seal heat calculations.

A pump is equipped with a single mechanical seal, which is operating at 3550 rpm and pumping water at 400 psig. The seal dimensions are as follows:

OD = 2.625 inches

ID = 2.125 inches

BD = 2.225 inches

Fsp = 42 inch pounds

Dp = 400 psi

N = 3550 rpm

f = 0.07

K = 0.5




The seal is pressurized at the outer diameter

1. The seal face area is determined as follows::

( ) in 1.862.1252.6254A 222 =−×

=π

2. The seal balance ratio is determined as follows:

0.8172.1252.6252.2252.625B 22

22

=−−

=

3. The seal spring force is determined as follows:

psi 22.61.8642

Psp ==

4. The total face pressure in pounds per square inch is determined as follows:

( ) psi 149.422.60.50.817400Ptot =+−×=

5. The mean face diameter in inches is calculated as follows:

( ) inches 2.3752

2.1252.625MD =+

=

6. The seal running torque is calculated as follows:

pounds-inch 23.0892

2.3750.071.86149.4RT =×××=

7. The seal power (hp) is calculated as follows:

( ) hp 1.3063,000

355023.089hp =×

=

8. The heat generated by the seal is calculated as follows:

BTU/hr 3311=25451.30Qheat ×=




If the temperature from the heat generated at the seal faces is high enough or if the liquid vapor pressure is high, the fluid film between the seal faces can vaporize. If the fluid film vaporizes, the seal faces will open and momentarily tilt as gas builds up and then escapes. Cool fluid will enter the seal and allow the faces to close once again. Frictional heat will cause local temperatures to rise once again, and the cycle is repeated. The cycle results in an unstable seal. The edges of the mating ring face will chip, and, in extreme cases, the mating ring face will pit and gross leakage will occur.

To prevent overheating the mechanical seal faces, a cooling fluid is often supplied to the seal chamber area. Cooling fluid is often called flush. Flush is a small amount of fluid that is introduced into the seal chamber on the process fluid side of the mechanical seal, in close proximity to the sealing faces. The flush provides cooling and lubrication for the seal faces. Flush is also used if the pumped fluid contains abrasives. Often, flush fluid is taken from the process stream and supplied to the mechanical seal after being cooled and cleaned (through the use of a filter or cyclone separator). If the abrasive levels of the process fluid are too high, or if the particulate in the process fluid is difficult to remove because of size, an external flush source can be used.




The flow rate of flush fluid can be determined from calculations shown in Appendix E.2 in API Standard 682. To determine the flush system flow rate, the maximum allowable temperature rise would be calculated by subtracting the maximum allowable temperature in the seal chamber from the flush fluid injection temperature. API Standard 682 recommends a maximum temperature of 5°F to 10°F. The following equation can then be used to determine the minimum flush fluid flow rate:

( )( )( )CdTMax SG500QQ heat

inj =

Where:

Qinj = Flush fluid flow rate

Max dT = Maximum differential temperature in degrees Fahrenheit

Qheat = Heat generation at the seal faces

SG = Specific gravity of the flush fluid

C = Specific heat of flush fluid at pump temperature (BTU/lbs.°F)

For this example, a maximum differential temperature rise of 10°F, a flush fluid specific gravity of 0.90, and a specific heat of 0.62 BTU/lbs.°F are assumed. Using the heat generated at the seal faces in the previous example (3311 BTU/hr), the following example illustrates the determination of the minimum flush flow rate:

( )( )( ) gpm 1.1870.62100.90500

3311Qinj ==

The minimum flush flow rate would be 1.187 gpm. API recommends a design factor of two, so the minimum flush flow rate should be 2.374 gpm.




In accordance with API Standard 682, only forced flush systems can be used. A forced flush system uses pressure source to circulate fluid through the seal chamber. A forced flush system ensures that flush fluid is circulating through the seal chamber. Thermosyphon (also called natural circulation or thermal head circulation) systems are not acceptable because there is no positive force to ensure that flush fluid is circulating through the seal during pump operation. The pressure source for the flush system can be the pump discharge pressure (orificed), an externally mounted pump, or a device internal to the mechanical seal, such as a pumping ring. In accordance with API Standard 682, if the mechanical seal is supplied with an internal pumping ring, a thermosyphon system should also be used to maintain flush fluid flow when the pump is idle.

Seal flush systems that are supplied from an external source must be equipped with indicators for seal chamber pressure and flush pressure. The pressure indicators provide a means of verifying that the flush system is operating properly.

API Standard 682, Appendix C, provides standard flush system configurations. Figure 34 shows examples of the flush system installations using API 682 plans.




Figure 34. Flush Systems Using API 682 Plans

In accordance with 31-SAMSS-012, seal flush connections for all seal types and porting arrangements must direct the flush fluid at the seal faces, maximize the uniformity of cooling fluid on the seal faces, and minimize flush fluid impingement on the faces or metal parts. Impingement on the seal faces should be minimized to prevent the velocity of the flush from eroding the sealing faces. Impingement is also minimized to prevent damage to metal components and seal faces from debris or particulate that may enter the flush system. All radial flush arrangements must be a multiport design to ensure adequate distribution of the flush fluid in the seal chamber. Single port flush connections must be tangential in the direction of shaft




rotation and must allow for complete venting of the seal chamber.

In addition to flush systems, mechanical seals can also be equipped with a quench system, a buffer system, or a barrier system. A quench system uses a neutral fluid directed at the atmosphere side of the seal to prevent or limit the formation of solids that may interfere with seal movement. Quenching is commonly used on mechanical seals when the pumped fluid is hydrocarbon and above 300 °F or other fluids that will crystallize when dried, such as caustic.

A buffer system uses fluid as a coolant and lubricant, or buffer, between dual mechanical seals. Buffer fluid is almost always at atmospheric pressure.

A barrier system uses fluid introduced between dual mechanical seals to completely isolate the pump process fluid from the environment. Barrier fluid pressure is always higher than the process pressure.

Saudi Aramco Standard 31-SAMSS-012 places the responsibility of selecting a compatible barrier or buffer fluid on the pump and seal system manufacturers. Some common fluids used as barrier or buffer fluids are low viscosity grade oil (such as grade 32), automatic transmission fluid (ATF), ethylene glycol (not automotive antifreeze), and water.

Seal Types - Regardless of the seal manufacturer, API 682 specifies three arrangements (arrangement 1, 2, and 3) of mechanical seals with several mechanical seal types in the arrangements. All mechanical seals designed to API Standard 682 are cartridge seals. A cartridge seal is a completely self-contained unit that includes all seal parts and that is preassembled and preset before installation. The arrangement 1 seal is a single, inside-mounted, balanced-type cartridge seal. The standard arrangement 1 seal is a Type A pusher seal. Other optional arrangement 1 seals are the Type B, which is a single, low-temperature, nonpusher (rotating bellows) seal, and Type C, which is a single, high-temperature, nonpusher (stationary bellows) seal. Figure 35 shows vendor representations of the variations of the arrangement 1 seals. Data sheet selections show where the seals can be reversed with respect to the flexible element being rotating or stationary.




Figure 35. Vendor Representations of Arrangement 1 Mechanical Seals




Arrangement 2 seals are dual, unpressurized, balanced, cartridge mechanical seals. The seal type can be either Type A, Type B, or Type C seals. Figure 36 shows a vendor representation of a Type A, arrangement 2 seal.

Figure 36. Vendor Representation of a Type A, Arrangement 2 Mechanical Seal

The inner seal of arrangement 2 mechanical seals must be designed with a positive means of retaining the sealing components and sufficient closing force to prevent the faces opening to a pressurization of the buffer fluid to 40 psig (2.75 bar). The outer seal must be designed to the same operating pressure as the inner seal. Cooling for the inboard seal is through the use of a flush system. The outer seal is cooled through the use of the circulating buffer fluid.




Arrangement 3 seals are dual-pressurized, inside, balanced, cartridge mechanical seals. The inner seal is designed with an internal (reverse) balance feature designed and constructed to withstand reverse pressure differentials without opening: this design feature is called “reverse pressure” capability. The internal or reverse balance feature requires that the mating ring and the secondary seal be designed to stay in place in the event that barrier pressure is lost. The seal will stay closed with internal pressure on the seal. If barrier pressure is lost, the arrangement 3 seal will operate as an unpressurized arrangement 2 seal. Barrier pressure should be maintained 20 psi to 60 psi (1.4 to 4.1 bar) over the pressure in the seal chamber. If the barrier pressure is too low, the seal will operate as an unpressurized arrangement 2 seal. If the barrier pressure is too high, the seal will overheat and fail. The arrangement 3 seal can use either Type A, Type B, or Type C seals. Figure 37 shows a vendor representation of a Type A, arrangement 3 seal.

Figure 37. Vendor Representation of a Type A, Arrangement 3 Mechanical Seal




Service Categories - API 682 defines three service categories that relate to mechanical seal operating conditions: flashing hydrocarbon service, nonflashing hydrocarbon service, and nonhydrocarbon service. Hydrocarbon service is defined as material that contains predominantly all hydrogen and carbon atoms; however, nonhydrocarbon constituents may be entrained in the process stream. Flashing hydrocarbon service is defined as any hydrocarbon service that requires vapor suppression by cooling or pressurizing to prevent flashing. Flashing hydrocarbon service includes all hydrocarbon services in which the fluid has a vapor pressure greater than 14.5 psia (1 bar) at pumping temperatures. Nonflashing hydrocarbon service is defined as all hydrocarbon services that will not require vapor suppression to prevent transformation from a liquid phase to a vapor phase. Nonflashing hydrocarbons, as defined by API 682, have a vapor pressure less than 14.5 psia (1 bar) at pumping temperature. Nonhydrocarbon service includes all services that cannot be defined as containing all hydrogen and carbon atoms; however, some hydrocarbons may be entrained in the process stream. Nonhydrocarbon service includes boiler feedwater and other water services, sour water, caustics, acids, amines, and other chemicals commonly used in refinery service. Service categories are used to identify the arrangement and type of seal used for an application, as well as the specific materials of construction for the parts of the mechanical seal. API Standard 682 provides seal selection and material guidance.




Lift-Off Face Seals

An alternate type of mechanical seal is the lift-off seal, which is also called a noncontact seal. The operation of the lift-off mechanical seal is similar to the operation of a contact mechanical seal, with the exception of the hydrodynamic forces between the seal faces. During operation, a lift-off seal generates a liquid or gas pressure between the seal faces, separating the seal faces slightly. When the pump is idle, spring pressure holds the seal faces together, forming a seal when the pump is not operating. Lift-off seals can be used for sealing of high vapor pressure (volatile) fluids in chemical and petrochemical processes. Lift-off seals can be designed as single or double seals. Figure 38 shows an example of a single, dry-gas, lift-off seal.

Figure 38. Single, Dry-Gas, Lift-Off Seal




Lift-off face seals are classified as pressurized, noncontact face seals. Lift-off face seals can be a dry or wet type. The lift-off seal has a rotating primary ring that is similar to the primary ring of a contact mechanical seal. The major difference between these two seal types is that the primary ring of the lift-off seal has a broader face with grooves that are cut into one face of the primary ring. The depth of the grooves is typically 0.0001 to 0.0002 inches. The groove design pattern varies with the seal manufacturer. Common patterns include uni-directional V-grooves cut in a spiral pattern, or bidirectional T-grooves.

The lift-off seal is supplied with a gas (vaporized process liquid, nitrogen, air, or steam) or a liquid supply. In this case, a geometry change is made to the seal face. Spiral grooves are incorporated into one of the seal faces to generate hydrodynamic lift to separate the seal faces. In the case of liquid buffer lift-off seals, this concept is used to move a small quality of buffer liquid from the low pressure to the high-pressure process liquid side of the seal. During pump operation, the groove patterns on the primary ring pumps gas or liquid between the primary and mating ring. As shown in Figure 38, pumping gas or liquid between the sliding faces raises the pressure of the gas or liquid that is in the pressure field between the faces, which forms a pressure dam that serves as a barrier to pump fluid flow. The rise in the pressure also causes the sliding faces to separate slightly, which prevents any form of face contact. The face separation is typically .0001 to .0002 inches. Separation of the seal faces occurs at low rpm. Since the primary to mating ring seal face gap is controlled to extremely small values, seal leakage is maintained at a small value, typically below 1 standard cubic foot per minute.

Depending on the seal manufacturer and design (liquid or gas), some lift-off seals must be supplied with filtered gas or liquid with the pressure regulated at 20 to 30 psig over the pumped fluid pressure.




The seal design that is shown in Figure 39 is considered uni-directional because the function of the spiral seal face will only work when rotated in one direction. If one-directional seals are used on both ends of a horizontal pump, one seal is designed to rotate in the opposite direction of the other seal. One-directional gas seals may be damaged if the pump rotation is reversed by process fluid backflow through the pump. Bidirectional seal faces eliminate seal damage from pump rotation reversal and minimize the amount of spare seal faces to be maintained.

Figure 39. Lift-Off Seal Rotating Face V-Groove

Lift-off seals provide the following advantages:

• Minimal product leakage to the environment

• Low power consumption

• Low heat generation

• Long operational life (low wear)

The disadvantages of lift-off seals include:

• Require a clean (filtered) gas or liquid supply

• Require a gas or liquid filter and regulating system




Auxiliary Seals

An auxiliary seal is a mechanical seal, either contact or lift-off face design, installed in tandem with a primary mechanical seal. The purpose of an auxiliary seal is to control pump leakage in the event that the primary seal fails. Auxiliary seals can be dry running or wet running, and they can be either contact or lift-off type. The auxiliary seal must be suitable for operation at full seal chamber conditions, but it normally operates dry or wetted below the seal leakage pressure switch setting. Figure 40 shows the configuration of a wet running auxiliary seal.

Figure 40. Wet Running Auxiliary Seal




Figure 41 shows the configuration of a dry running auxiliary seal.

Figure 41. Dry Running Auxiliary Seal

In the event of primary seal failure, the auxiliary seal will minimize the leakage until the pump can be isolated and shut down. The primary seal operation is monitored by sensing the pressure at the seal outlet for the barrier or buffer system. If the primary seal fails, the increase in pressure from leakage past the primary seal will activate an alarm. In high pressure applications, such as large natural-gas liquids (NGL) injection pumps, the high barrier or buffer system return pressure will activate an emergency shutdown of the pump.

In accordance with 31-SAMSS-012, auxiliary dry-running seals must be face-type seals, that are capable of 25,000 hours of operation in a gas or liquid environment at a pressure up to the setpoint of the seal leakage detection pressure switch and 1,000 hours of operation in a liquid environment up to the maximum seal chamber conditions.




PUMP BEARING ARRANGEMENTS AND LUBRICATION REQUIREMENTS

The function of bearings in centrifugal pumps is to keep the shaft or rotor in correct alignment with the stationary parts under the action of radial and transverse loads. Bearings that maintain radial shaft positioning are commonly referred to as line or journal bearings. Bearings that maintain the axial shaft position are called thrust bearings. The types of bearings used in centrifugal pumps are described in MEX 211.01. This section of the module describes the bearing arrangements commonly used in centrifugal pumps with regard to bearing loads.

Bearing Loading Bearing load is defined as the force in pounds or newtons that is felt on the bearing during operation. Radial load is the force on a bearing radially from the shaft. Axial load or thrust is the force on a bearing axially along the shaft.

Radial Loads

Some single-volute pump casings are designed for uniform or near uniform pressures (specific speed range between 500 and 3500 English units) around the volute casing at the BEP. The uniform pressures act equally on all sides of the impeller when the pump is operated at design capacity.




Figure 42A illustrates the equal pressure that surrounds a pump impeller during operation at design capacity. At capacities other than design capacity, the pressures around the impeller are not uniform and there is a resultant radial load (thrust). Figure 42B illustrates the unequal pressure that surrounds the pump impeller and the resultant force (F) when the pump is operating at capacities other than the design capacity.

Figure 42. Radial Force

The magnitude of the resultant radial force on the impeller during operation at capacities other than the design capacity can be calculated through use of the following equation:

))(b)(SG)(H)(DKk(F 22rr =

Where:

Fr = Radial force

k = 0.433 (9790 metric)

Kr = Radial force factor coefficient

SG = The specific gravity of the fluid pumped

H = Pump head in feet or meters

D2 = Outside diameter of the impeller

b2 = Width of the impeller at the discharge, including shrouds




The radial force factor coefficient has been determined experimentally as a function of specific speed and capacity. Figure 43 shows a graph of specific speed (English units) versus radial force coefficient. The three plots are for when the pump is operating at or near the capacity at BEP, when the capacity is one half the capacity at BEP, and when the pump is operating at shutoff head.

Figure 43. Radial Force Factor Coefficient Plot

The following example illustrates the use of the resultant radial load calculations to determine the magnitude of the resultant radial force. A single-volute pump is operating with a specific speed of 2000 while pumping cold water (SG = 1.0) at a shutoff head of 252 feet. The pump impeller diameter is 15.125”, and the width of the impeller is 2.5”. Kr is determined from the graph in Figure 43 for shutoff head with a value of approximately 0.31.




The resultant radial thrust can be calculated through use of the following equation:

))(b)(SG)(H)(DKk(F 22rr =

pounds 1279=2.5))(15.125)((1.00)(252.433(0.31)Fr =

Radial load can be lowered throughout the entire capacity range through the use of a double-volute or a diffuser-type casing. The use of a double-volute or diffuser-type casing should be considered when a pump is normally operated at variable capacities, especially at shutoff head.

As described in the casing designs section in MEX 211.01, radial load is minimized in multistage centrifugal pumps by staggering the volutes and by canceling out opposing radial thrusts.

Axial Loads

Axial load is the sum of the unbalanced forces that act on the impeller in the axial direction (axial thrust) and, in the case of vertical pumps, the force and the mass force of the pump rotor assembly. Axial pump loads vary with the type of pump and impeller. Figure 44 shows the resultant axial thrust for a horizontal, single-stage, single-suction, closed-impeller pump.

Figure 44. Hydraulic Axial Thrust Produced by a Horizontal, Single-Stage, Single-Suction, Closed-Impeller Pump




For overhung, single-stage pumps, an additional axial force, which results from the difference between atmospheric pressure and suction pressure, is felt on the shaft area. Figure 45 illustrates this additional axial force. This force acts towards the impeller when the suction pressure is less than atmospheric, and it acts in the opposite direction when suction pressure is higher than atmospheric.

Figure 45. Additional Axial Thrust on an Overhung Pump

When the single-suction closed impeller is used in multistage pumps, the net hydraulic thrust is the sum of the axial thrust produced by each impeller in the pump.




The difference in pressure between the suction and the discharge of the pump acting on the area of the pump impeller suction provides an axial thrust on the impeller. Double-suction pumps are often used to minimize the effect of axial thrust from an impeller. Figure 46 shows the resultant hydraulic force for a horizontal, single-stage, double suction, closed-impeller pump.

Figure 46. Hydraulic Axial Thrust Produced by a Horizontal, Single-Stage, Double-Suction, Closed-Impeller Pump




In practice, the hydraulic balance on a double-suction pump may not be achieved for the following reasons:

• The suction passages to the two suction eyes may not provide equal or uniform flows to the two sides.

• External piping conditions, such as an elbow located too close to the pump suction nozzle, may cause unequal flow to the two suction eyes. Proper piping arrangements to double suction impellers are of utmost importance. As a rule, three to five straight pipe diameters must be present downstream of an elbow to ensure that equal flow enters each side of the impeller, and the suction pipe should be perpendicular to the pump axis.

• The two sides of the discharge casing waterways may not be symmetrical, or the impeller may be located off-center. These conditions will alter the flow characteristics between the impeller shrouds and the casing, and thereby cause unequal pressures on the shrouds.

• Unequal leaking through the wear rings on either side of the impeller can result in differential pressure across the impeller and upset the balance.




Axial thrust in horizontal, single-suction, semi-open, radial flow impellers is illustrated in Figure 47. The pressure on the open side of the impeller varies from essentially the discharge pressure at the periphery (diameter D2) to the suction pressure at the impeller eye (diameter D1). The pressure distribution at the back of the impeller shroud varies from discharge pressure at the periphery to a slightly lower pressure at the impeller hub. The unbalanced portion of the axial thrust on the impeller is represented by the crosshatched area in Figure 47.

Figure 47. Axial Thrust in Horizontal, Single-Suction, Semi-Open, Radial Flow Impellers

Axial loading for vertical pumps must take the weight of the rotor assembly (shafting, couplings, and impellers) into consideration when determining the axial load.

Thrust Direction

The sum of the axial loads in one direction is balanced against the sum of the axial load in the opposite direction, and this situation results in a net active thrust in one direction. Active thrust is defined as the normal thrust direction when a pump is operating. The axial direction opposite the direction of active thrust is called the direction of inactive thrust. Depending on the pump design, the direction of active thrust can change with changes in pump capacity.




The direction of active thrust can cause tension or compression in the shaft. Shaft tensile stresses can occur when the direction of active thrust is away from the thrust bearing. Compressive stresses can occur when the direction of active thrust is towards the thrust bearing. The direction of axial thrust varies with flow rate, which results in a compressive or tensile stress on the pump shaft. The magnitude of thrust varies with the pump design. Figure 48 shows examples of compressive and tensile stresses on pump shafts for horizontal and vertical pumps. If a pump is operated in the discharge recirculation zone, the stresses on the pump shaft can cycle between compressive and tensile. Repeated cycling between compressive and tensile stresses can cause pump thrust bearing damage and shaft damage from high axial loads and from fatigue cracking corrosion.

Figure 48. Examples of Directions of Axial Thrust




Thrust Balancing Designs

The amount of active axial thrust in single-stage and multiple-stage pumps can be minimized through the use of the following different methods:

• Back and front wear rings with impeller balance holes

• Pumpout vanes

• Double suction impellers

• Stacked impeller design

• Opposed impeller design

• Balance drum




Back and Front Wear Rings with Balance Holes - The ordinary, single-suction, closed, radial impeller with the shaft passing through the impeller eye is subject to axial thrust because a portion of the front impeller wall is exposed to the suction pressure while the area in back of the impeller wall is exposed to the discharge pressure. If the discharge chamber pressure was uniform over the entire impeller surface, the axial force acting toward the suction would be equal to the product of the net pressure generated by the impeller and the unbalanced annular area. In actual use, the pressure on the two single-suction closed impeller walls is not uniform. The liquid trapped between the impeller shrouds and the casing walls is in rotation, and the pressure at the impeller periphery is higher than at the impeller hub. Figure 49 illustrates the actual pressure distribution across an impeller.

Figure 49. Actual Pressure Distribution across an Impeller




To minimize the axial thrust of a single-suction impeller, the pump impeller can be equipped with both front and back wear rings. The front and back wear rings effectively isolate the high pressure and low pressure areas of the impeller. The thrust areas are equalized through the use of the same inner diameter of both the front and back wear rings. Pressure that is approximately equal to the suction pressure is maintained in a chamber located on the impeller side of the back wear ring by the drilling of balance holes through the impeller. Figure 50 shows an example of a single-suction impeller equipped with front and back wear rings and balance holes.

Figure 50. Front and Back Wear Rings and Balance Holes

Leakage past the back wear ring is recirculated back to the pump suction through the balance holes. Large (greater than 10” suction), single-stage, single-suction pumps do not commonly use balance holes because the leakage from the back wear rings through the balance holes opposes fluid flow through the suction of the impeller and creates disturbances that can affect the pump capacity. Large, single-stage, single-suction pumps commonly use a piped connection from the area behind the impeller to the pump suction piping.




Pump-Out Vanes - The primary function of pump-out vanes is to minimize packing or seal leakage by reducing the fluid pressure on the seal chamber. Pump-out vanes also prevent foreign material that can be suspended in the pumped fluid from lodging in the clearance space between the shroud and the adjacent wall of the casing. Reducing pressure behind the impeller shroud with pump-out will also reduce axial thrust. Figure 51 illustrates the effect of pump-out vanes on the pressure differential across an impeller.

Figure 51. Pressure Differential across an Impeller with Pump-Out Vanes

Double Suction Impellers - As mentioned previously, double-suction impellers can be used to minimize the axial load from hydraulic thrust.




Stacked Impeller Design - The stacked impeller design is used on multi-stage pumps. The stacked design consists of several single-suction impellers mounted on one shaft, each having its suction inlet facing in the same direction and its stages following one another in ascending order of pressure. Thrust increases with the increasing number of impellers in the stacked impeller; however, the stacked impeller design axial thrust is balanced by a single hydraulic balancing device (a balance drum, which is discussed later in this section). Figure 52 shows an example of a multi-stage pump using the stacked impeller design and a hydraulic balancing device.

Figure 52. Stacked Impeller Design with Hydraulic Balancing Device




Opposed Impeller Design - The opposed impeller design is used on multi-stage pumps. The opposed impeller design consists of single-suction impellers mounted on a single shaft, with a portion of the impellers facing one direction and the other impellers facing the opposite direction. With this arrangement, axial hydraulic thrust is minimized by balancing the thrust of one group of impellers against the opposite group of impellers. When an even number of impellers is used, typically one-half of the impellers face one direction, and the other half of the impellers face the opposite direction. When an odd number of impellers is used, the pump shaft diameter and the interstage bushing diameters are varied to provide the effect of a hydraulic balancing device that will compensate for the hydraulic thrust on one of the stages. Figure 53 shows an example of a multistage pump that uses the opposed impeller design.

Figure 53. Opposed Impeller Design




Balance Drum - A balance drum, which is also known as a balance piston, is a hydraulic thrust-balancing device used to reduce the axial thrust in a pump. There are two types of devices that are commonly used to balance axial thrust in centrifugal pumps: a balance drum and a balance disk.

A balance drum is shown in Figure 54. The balancing drum is either keyed or screwed to the pump shaft and separates the balancing chamber at the back of the impeller (or if multi-staged, an end-stage impeller) and the interior of the pump. A balancing drumhead is fixed to the pump casing, and it allows for a small radial clearance that separates the drum and the stationary portion of the balancing device.

Figure 54. Balancing Drum

The area on the seal chamber side of the balance drum is subjected to the pump suction pressure. The area on the impeller side of the balance drum is exposed to the high-pressure fluid in the pump. The difference in fluid pressure across the balance drum provides a force on the balance drum that is opposite to the direction of axial hydraulic thrust from the impellers. The typical balance design is 90 to 95 percent of total axial impeller thrust. Any residual thrust that is not balanced by




the balance drum is absorbed by the thrust bearing on the end of the shaft. The amount of residual thrust that must absorbed by the thrust bearing changes as a function of the differential pressure from the suction pressure and the internal pump pressure. The use of a balance drum enables the selection of a smaller thrust bearing, which results in lower horsepower losses.

Another form of balancing device is called a balancing disk. Similar to the balancing drum, the balancing disk also uses a balancing chamber. The balancing disk is secured to the shaft, and the balancing disk head is fixed to the casing, as shown in Figure 55. The leakage to the balancing chamber flows through a small axial clearance between the balancing disk and the balancing disk head. The liquid, depending on system design, would then flow to either the pump suction or back to a tank. A restricting orifice is typically placed in the leakage return line. The orifice provides backpressure in the balance chamber by restricting fluid flow out of the balance chamber. The balance chamber backpressure is required for the proper operation of the balancing disk.




Figure 55. Balancing Disk




The balance disk automatically compensates for changes in axial impeller thrust by varying the amount of axial clearance between the balancing disk and the balancing disk head. For example, if the impeller thrust increases, the disk moves towards the disk head and reduces the clearance between the disk and the disk head. The reduction in clearance reduces the amount of leakage from the impeller side of the disk to the balance chamber. The reduction of leakage to the balance chamber reduces the backpressure in the balancing chamber. This drop in pressure provides a higher differential pressure, from the discharge pressure side to the balance chamber, across the balance disk. The higher pressure on the discharge pressure side of the balance disk provides the force to oppose the axial hydraulic thrust from the impeller(s), and it allows the disk to move away from the disk head until a balanced axial thrust equilibrium is achieved.




Bearing Types The functions of bearings in a pump are to support the rotor, to minimize friction losses, and to position the rotor with respect to the stationary pump components. Because of the close clearances between the pump impeller and the casing, the position of the rotor is extremely important in pump operation.

Two types of bearings are used on pumps: antifriction bearings and hydrodynamic bearings. This section briefly discusses the common types of antifriction and hydrodynamic bearings that are used on pumps in Saudi Aramco applications.

Centrifugal pumps are fitted with bearings that are appropriate for the application and pump design. Figure 56 lists the types of bearings that are commonly used for the different pump applications.

Figure 56. Typical Pump Bearing Applications




Antifriction

Antifriction bearings, which are so called because they are designed to overcome friction, are of two types: ball bearings and roller bearings. In a hydrodynamic bearing, the frictional resistance to sliding motion is overcome by separating the surfaces with a fluid film. The antifriction bearing substitutes rolling motion for sliding motion through the use of rolling elements between the rotating and stationary surfaces, and this rolling motion reduces friction to a fraction of the friction that exists where hydrodynamic bearings are used.

Basically, all antifriction bearings consist of two hardened rings that are called the inner and the outer rings or races, the hardened rolling elements that may be either balls or rollers, and a separator or cage. Bearing size is usually given in terms of boundary dimensions. Boundary dimensions are the outside diameter, the bore, and the width. The inner and outer rings provide continuous tracks or races in which the rollers or balls roll. The separator or retainer properly spaces the rolling elements around the track and guides the rolling elements through the load zone. Other words and terms that are used in describing antifriction ball bearings are the face, the shoulders, or the corners. The terms that are used to describe taper-roller bearings are a little different in that the outer ring is called the cup, and the inner ring is called the cone. The word “cage,” rather than separator or retainer, is standard for taper-roller bearings.

Classified by function, ball bearings used in pumps may be divided into two groups: radial and thrust bearings (angular-contact bearings). Angular-contact bearings can support combined radial and thrust loads.

Radial - Radial bearings are designed primarily to carry a load in a direction that is perpendicular to the axis of rotation. There are several different types of radial antifriction bearings, including single-row ball bearings and double-row ball bearings.




Single-row ball bearings are the most widely used of all ball bearings and probably of all antifriction bearings. Single-row ball bearings can sustain combined radial and thrust loads, or thrust loads alone if they are angular-contact type bearings, in either direction, even at extremely high speeds. In accordance with API Standard 610, all ball bearings must be Conrad-type bearings. Conrad-type bearings are named after Robert Conrad, who invented the Conrad bearing assembly method. The Conrad bearing assembly method does not use ball filling notches or slots.

A cross-section of the single-row ball bearing is illustrated in Figure 57. The ball element is positioned between an inner race and an outer race. Single-row, angular-contact ball bearings can have equal load-carrying capabilities in either direction or one direction only, depending on design. Such ball bearings are recommended for moderate to heavy radial loads, for thrust loads in either direction, or for combination loads.

Figure 57. Single-Row Ball Bearing




Another type of ball bearing that is used for high radial load is the double-row, angular-contact ball bearing shown in Figure 58. The double-row ball bearing uses the same principle of design as the single-row bearing. However, the grooves for the two rows of balls are positioned so the load through the balls tends to push outward on the outer ring races. The use of two rows of balls increases the radial load capacity of the bearing and provides an increase in the amount of thrust capacity.

Figure 58. Double-Row, Angular-Contact Ball Bearing




Thrust - Angular-contact bearings, as shown in Figure 59, can support radial loads when combined with thrust loads in one direction. The inner and outer races are made with an extra high shoulder on the thrust side. Angular contact bearings are designed for combination loads in which the thrust component is greater than the capacity of single-row ball bearings.

Figure 59. Angular-Contact Bearing




The double row spherical roller bearing, which is shown in Figure 60, is a self-aligning bearing that utilizes rolling elements that are shaped like barrels. The outer race has a single spherical raceway. The inner race has two spherical raceways that are separated by a center flange. The double row spherical roller bearing will support a heavy radial load and a heavy thrust load in both directions. Double row spherical roller bearings are inherently self-aligning because the assembly of the inner race, the center flange, and the rollers is free to swivel within the outer race.

Figure 60. Spherical Barrel-Shaped Roller Bearing




Hydrodynamic

Some centrifugal pumps are equipped with hydrodynamic bearings. Hydrodynamic bearings (also known as “self-acting” bearings) depend on the relative motion of the journal (the portion of the shaft that is supported by the bearing) to produce a fluid film pressure for load support. The term “hydrodynamic” refers to the relative motion of the journal in relation to the bearing that produces the fluid film pressure. A hydrodynamic bearing is different from a hydrostatic bearing. Hydrostatic bearings (also known as “externally pressurized” bearings) achieve load support by the supply of fluid from an external high-pressure source and require no relative motion of the journal.

Hydrodynamic bearings are generally very simple in construction and operation. These bearings are efficient, and they can support extremely heavy rotating loads. The ability to carry the heavy loads is due to an oil wedge that is developed between the journal and the bearing surface. The surface of the shaft journal slides over the surface of the bearing. A converging wedge fluid film is automatically generated in a lubricated journal bearing by the running clearance between the journal and the bearing bore, combined with the effect of load and rotation, which produces a displaced, eccentric disposition of the journal. The clearance is generally 0.0015 per inch of shaft diameter; however, the appropriate bearing instruction book must be consulted for actual bearing clearances. The thickness of the oil wedge at the maximum load point is usually 15 to 25 microns. Figure 61 shows the formation of the oil wedge during shaft rotation. When the shaft rotates, the fluid tangential force and the fluid radial force act against the bearing load to lift the journal up and push the journal to the side.




Hydrodynamic bearings offer great rigidity and stability under extreme load conditions. Journal bearings are much less sensitive than roller bearings to vibration, poor fits, corrosion, and contaminants. The fluid film within a hydrodynamic radial bearing absorbs (viscous dampening) much of the force exerted by radial vibration of the journal within the clearance between the shaft and the bearing’s Babbitt lining (discussed below).

Figure 61. Shaft/Bearing Dynamics




Hydrodynamic flow is controlled by the bearing dimensions, the geometry, the clearance, the journal's eccentricity, and the shaft running speed. The heat generation or power consumption in a bearing film is produced by the work that is done in shearing the film. The bearing material must be of adequate hardness and strength to support the load. Conversely the material must soft enough to embed particulate debris to prevent damage to the pump shaft. Good contact compatibility is important. If lubrication becomes marginal, the material must be able to withstand momentary contact with the shaft. Corrosion resistance is necessary to prevent the acidic products of oxidation in the lubricating oil from corroding the bearing surfaces. The bearing material must satisfy all of these requirements.

In general, the softer bearing materials (e.g., the whitemetals or babbitt) are best for high-speed lightly loaded pump applications, and they will withstand occasional contact without serious damage. Babbitt or whitemetal thicknesses vary from 0.20” to as little as 0.050” based on bearing application. For highly loaded applications, a thin coating of a heat-conductive material (copper or bronze) is placed between the whitemetal and steel layers. These bearings are typically called “tri-metal” bearings.

When classified by function, hydrodynamic bearings may be divided into two main groups: radial and thrust.

Radial - A typical plain journal bearing is shown in Figure 62. The bearing is made in two halves that are split at the horizontal centerline. The bearing is commonly called a split sleeve bearing, which is made in two halves that are split at the horizontal centerline. The split halves are bolted and doweled together to ensure proper alignment and support for both halves. Other types of sleeve bearings may be constructed from a solid piece. The bearing liner has a Babbitt bore that acts as the bearing surface. Babbitt is a tin-lead material that serves as a soft sacrificial lining and that has a melting temperature of approximately 300°F. The outer diameter of the bearing liner is either a flat or spherical shape (self-aligning) that mates with a similar shape in the supporting pedestal. The spherical shape assists to lock the bearing liner in place and accommodates minor internal misalignment between a set of radial bearings. Oil passages are drilled through the pedestal and bearing liner to provide an oil inlet to the oil grooves of the bearing.




An oil film separates the shaft journal from the bearing babbitt surface. Oil films may vary from .0006 to about .001 inch (15 to 25 microns) in thickness at the point of minimum thickness. The oil film thickness generally depends upon the load, the viscosity of the oil, the shaft speed and, to some extent, the bearing clearance.

Figure 62. Typical Journal Bearing

Another type of journal bearing that is used in pumps is the multi-lobe journal bearing, which is shown in Figure 63. The multi-lobe bearing has a series of slight depressions that run the length of the bearing parallel to the axis of the shaft. The depressions provide the bearing with the appearance of a series of lobes. The number of lobes varies from two to four over the circumference of the bearing.




The multi-lobe bearing that is shown in Figure 63 has equal spacing between the depressions. The spacing is off centered from the bottom of the bearing to prevent the journal from resting in a depression. The equal spacing produces an equal lobe pattern around the bearing that breaks up the hydraulic film and produces better stability during operation. At shaft speeds above approximately 6,000 rpm, failure to break up the hydraulic film will cause the fluid to develop a circumferential velocity. The force associated with this velocity will cause the journal to vibrate in a circular or elliptical orbit and at a frequency that is 47 to 53 percent (.47× to .53×) shaft rotative speed. Referred to as oil whirl, this pattern of vibration can cause damage to the bearing. An aggravation of oil whirl is referred to as oil whip. Oil whip occurs when the vibration frequency induced by oil whirl is the same as the natural vibration frequency of the rotor system; therefore, the vibration is resonant. The vibration is frequency and the consequently high displacement amplitude continue regardless of changes in the rotative speed. Turbine failure will occur within minutes or hours of inception of oil whip.

Figure 63. Multi-Lobe Journal Bearing




A third type of radial journal bearing that is used on pumps is called the pressure dam bearing or pressure bearing. Its construction is essentially the same as the split sleeve radial journal bearing with the exception that a shrouded step or dam is cut into the upper half of the bearing. The viscous and inertia effects combine to build up pressure over the top half of the bearing that, in effect, places an artificial load on the lower half of the bearing. The artificial load results in an increase in the eccentricity ratio, which increases the bearings stability and the bearing anti-whirl properties. It is important to note that the pressure dam must be in the upper half of the bearing. Installation of the pressure dam in the lower half of the bearing will cause bearing instability. Most pressure dam bearings are indexed to prevent improper assembly; however, older pressure dam bearing designs may not be indexed.

A common type of radial hydrodynamic bearing that is used in higher speed (greater than 4000 rpm) pumps is called the tilting-pad bearing, which is shown in Figure 64. The tilting-pad bearing is more stable than the standard journal bearing. Because of their excellent stability, tilting-pad bearings are usually used where shaft stability is a problem. The increased stability helps to provide a greater degree of damping of the turbine rotor, which avoids erratic vibration patterns. Tilting-pad bearings are used to prevent oil whirl.

The tilting-pad bearing is made up of a series of identical babbitted pads or shoes. The shaft journal rides on an oil film (0.0006” to 0.001”) and is supported by the pads. The number of pads that are used in a tilting-pad bearing can vary. The more common designs use four, five, or six pads. The bearing pads are made from a high-grade steel forging that is lined with a high-quality, tin-base Babbitt.

The pads usually have projections that fit under protection lips that are machined in the bearing casing to prevent radial movement of the pads. Circumferential movement is prevented by pins that pass through the bearing housing into holes in the pads.

To allow the pads to tilt freely without seizing or binding, the holes in the pads into which the pins fit are larger than the diameter of the pins. The tilting action of each pad takes place because the back half of the pad is machined to a radius that is less than the radius of the housing.




Oil enters the housing through oil inlets that are located between the pad segments. The oil discharges through drain orifices that are provided in the housing. The tilting-pad helps to form the oil wedge between the bearing surfaces and the shaft journal. Despite the relatively short span of the pads, the load capacity of tilting-pad bearings is similar to the load capacity of conventional bearings.

Figure 64. Typical Tilting-Pad Bearing

Thrust - Thrust bearings prevent axial (back and forth) movement of a pump rotor, which can cause the pump impeller to make contact with the pump casing. Such contact can cause extensive damage or stop (seize) the pump. Several types of hydrodynamic thrust bearings are used in pump applications, such as plain thrust bearings, tapered land thrust bearings, and tilt-pad thrust bearings. The load capacity of a plain thrust bearing is relatively low. Radial grooves are cut into the thrust




face as a passage for oil to cool and to lubricate the bearing surface. Tapered land thrust bearings resemble plain thrust bearings. Figure 65 shows a tapered land thrust bearing. The surface of the tapered land thrust bearing is divided into a number of pads. Each pad is separated by an oil feed groove. Each pad is tapered in a circumferential direction and radial direction. The taper allows the rotary motion to wipe oil into the contacting, wedge-shaped area, which builds up a load-carrying oil wedge. The exact form of the pad surface profile is not especially important. However, a flat land at the end of the tapered section is necessary to avoid excessive local contact stress under start-up conditions. The land should extend across the entire radial width of the pad, and it should occupy about 20 percent of the pad circumferential length.

Figure 65. Tapered Land Thrust Bearing




The tilting-pad thrust bearing, as shown in Figure 66, is the most satisfactory thrust bearing for heavy axial thrust, regardless of whether the shaft speed is high or low, or whether the shaft is horizontally mounted or vertically mounted. Because the pads are pivotally supported and are able to assume a small angle relative to the moving collar surface, the tilting-pad bearing is able to accommodate a large range of speed, load, and viscosity conditions. The pivotal movement and the small angle enable a full hydrodynamic fluid film to be maintained between the surfaces of the pad and the collar. The tilting-pad thrust bearing has the inherent advantage of being able to absorb significant amounts of thrust without excessive horsepower losses.

Figure 66. Tilting-Pad Thrust Bearing




The self-equalizing thrust bearing shown in Figure 67 is an alternative design of the tilting-pad thrust bearing. The self-equalizing thrust bearing is one of the most common types of tilting-pad thrust bearings that are used on rotating machinery.

The self-equalizing thrust bearing consists of a thrust collar, stationary shoes, leveling plates, and a base ring. The thrust collar can be either a separate component that is mounted on the turbine shaft, or it can be machined as an integral part of the turbine shaft. When the shaft is stationary, the stationary shoes lie with their surfaces parallel to the thrust collar. As the thrust collar starts to rotate, an oil film is created between the stationary shoes and the thrust collar. As the oil film builds up, each stationary shoe tilts to an angle that generates the proper distribution of the oil film pressure. The tilt of each shoes creates an oil wedge under each shoe. The oil wedge film can carry a heavier load than the flat film of oil for any given size.

Any axial movement that is generated by the pump shaft is transmitted through the stationary shoes to the leveling plates. The leveling plates uniformly distribute the thrust load around the bearing to equalize the axial thrust that is generated by the shaft. A misaligned load will force the leveling plates at the heaviest load point to tilt, which pushes the remaining stationary pads outward to equalize the load over all of the stationary pads. The leveling plates transmit the axial thrust to the base ring.

Similar to the load bearing section of a radial hydrodynamic bearing, thrust bearing shoe surfaces may use the tri-metal design for high load applications.

Self-equalizing thrust bearings are used as both double acting thrust bearings and single acting thrust bearings. Double-acting thrust bearings have stationary shoes and leveling plates on both sides of the thrust collar. Double-acting thrust bearings are used in applications in which axial thrust can be exerted in both directions. Single-acting thrust bearings have stationary shoes and leveling plates only on one side. Single-acting thrust bearings are used in applications in which axial thrust can be exerted in only one direction.




Figure 67. Self-Equalizing Thrust Bearing

Pump Industry Standard Bearing Applications

Understanding centrifugal pump dynamics and operating characteristics is an important part of extending pump life. During operation, hydraulic conditions in centrifugal pumps generate major axial thrust loads. The type and size of a pump’s impeller determine the load’s magnitude and duration. Steady radial loads are imposed by the weight of parts and components. Fluctuating radial loads result from hydraulic and unbalanced conditions of the rotor assembly.




Most centrifugal pumps are designed to either API Standard 610 or ANSI Standard B73.1. Heavy-duty process pumps are typically designed to API standards, while light-and medium-duty pumps are typically designed to ANSI standards.

The bearing arrangements for API and ANSI configurations vary slightly. In API designs, there are bearings at two positions: a single-row deep-groove ball bearing near the impeller, and an angular contact ball bearing set in the thrust position at the coupling end. Figure 68 shows a typical API bearing configuration.

Figure 68. Typical API Bearing Configuration




The bearing accommodates radial loads and floats in the housing. The bearing arrangement allows for thermal expansion of the shaft. The angular contact set is fixed in the housing. The angular contact set secures the impeller in the proper axial position, which handles the thrust load and a portion of the radial load.

ANSI pumps are very similar in bearing configuration to the API pumps with one notable exception. Instead of an angular contact bearing set, ANSI pumps use a single or double row deep-groove ball bearing.

Both API and ANSI pumps use the same bearing configurations in vertical pumps as horizontal pumps.

In both API and ANSI bearing configurations, the bearings in the thrust position must handle a complex, dynamic combination of thrust and radial loads, and the bearings must hold the entire pump rotor assembly in place under varying conditions. Several bearing types have been used as thrust bearings. The duplex angular contact ball bearing design used in API pumps consists of two bearings mounted back-to-back. Typically, each bearing uses a 40° contact angle, which means that the balls in each bearing roll on an axis that forms a 40° angle with the perpendicular radial position. Bearings with high contact angles, such as the 40° angles used in the API configuration, are designed to carry axial loads; they are axially rigid and radially soft. Bearings with low contact angles are designed for radial loads; they are axially soft and radially rigid.

Some problems can occur with high contact angle bearings, such as ball skidding. During operation, thrust loads apply pressure against one bearing and cause the one bearing to deflect and the other bearing to unload. Centrifugal forces that act on the unloaded bearing may cause its balls to run on a skewed axis and begin to skid. Ball skidding produces a microscopic wear or lapping process that damages the bearing raceway. The friction from ball skidding generates heat that reduces the bearing lubricant’s operating viscosity.




One method used to minimize the effect of ball skidding in API bearing configuration is to use a set of preloaded, 40° angle, angular contact bearings. When the active bearing in the set is under load, the residual preload in the other bearing will prevent unloading. The amount of preload for the bearing set can be difficult to determine because the temperature differential between the inner and outer bearings and the impeller thrust load cannot be determined with any degree of certainty. Improper preload on a bearing can cause the bearing to run hot and shorten bearing life.

An alternate method of reducing the ball skidding problem is to use a matched pair of angular contact bearings that have different contact angles. A common method uses a 40° angle and a 15° angle angular contact bearing set. The 40° bearing handles the axial thrust, while the 15° bearing handles any reversing thrust during startup (mainly from caused by the driver) and handles radial load during normal operation.

ANSI pumps are also susceptible to the same problem of ball skidding; however, double row deep groove ball bearings are not presently manufactured with different contact angles.

API Standard 610 provides guidance on bearing selection. Antifriction bearings should have a L-10 rating life of at least 25,000 hours with continuous operation at rated conditions, and at least 16,000 hours at maximum radial and axial loads and rated speed. The L-10 rating for antifriction bearings is defined as the number of hours at rated bearing load that 90% of a group of identical bearings will complete (25,000 hours of continuous operation) or exceed before evidence of failure. Failure evidence is generally defined as a 100% increase in measured vibration.




L-10 life for all friction bearings can be calculated through the use of the following equation:

P

10 FC

N16700L

=

Where:

L10 = L-10 rating in hours

N = Pump speed

C = The total force required to fail the bearing after 1,000,000 revolutions

F = The total of all actual forces acting on the bearing

P = 3 for ball bearings and 10/3 for roller bearings

Small changes in force (F) acting on an antifriction bearing can significantly reduce the L-10 life of a bearing.

API Standard 610 also provides a limit on antifriction dN number (also called the dmN or Ndm number). The dN number is a measure of the amount of frictional heat generated by a bearing. The dN number is calculated by using the following equation:

(rpm) speed bearing )millmeters (in bore BearingdN ×=




API Standard 610 limits the dN rating to less than 500,000. The bearing dN number can also be used to determine the type of bearing used and the type of lubrication required for the bearing. Table 1 shows a table that lists bearing dN number ranges, bearing types, and lubrication types.

dN Range Bearing Type Lubrication Type

Below 100,000 Antifriction Grease, sealed

100,000 to 300,000 Antifriction Regreasable

Below 300,000 Antifriction Oil lubricated, nonpressurized

Above 300,000 Sleeve, multi-lobe, or tilt pad Oil lubricated, pressurized

Table 1. dN Rating/Bearing Type/Lubrication Type Table

API Standard 610 requires the use of hydrodynamic radial and thrust bearings when the product of pump rated power, kW (hp) and rated speed, (rpm), is 4.0 million (5.4 million) or greater.




Lubrication Requirements Bearing lubrication requirements vary with the types of bearings used and the type of pump. There are two methods of bearing lubrication for centrifugal pumps: grease lubrication and oil lubrication. API Standard 610 specifies that all bearings and bearing housings be arranged for hydrocarbon oil lubrication. Saudi Aramco Standard 31-SAMSS-004 requires that single casing volute lineshaft and volute cantilever pump thrust bearings must be designed for grease lubrication.

Grease Lubricated

Bearings can be lubricated with grease by a variety of methods. Most greased bearings used on pumps are packed with grease prior to assembly. Some bearing housings are equipped with a grease fitting that will allow periodic application of grease to the bearing to make up for grease that has leaked out of the housing. Bearing housings should never be packed more than half-full because too much grease can cause bearing overheating and can produce pressure that will cause leakage and damage to the grease seals.

Other forms of grease lubrication methods include grease cups. Grease cups can be manual or automatic. A grease cup can be threaded to the bearing housing, or it can be centrally located with lines running to various lubrication points. The grease cup contains a small quantity of grease. On a manual grease cup, a handle is periodically turned to admit a small amount of grease to the bearing. Automatic grease cups typically use a spring cap that maintains a slight positive pressure on the grease. For hard-to-reach lubrication points, some spring-loaded grease cups have lines leading from the cup to the point of application. A shortcoming of this method is the tendency of the grease to separate into oil and solid phases as a result of the constant pressure exerted by the spring on the small volume of grease in the system.

If several pumps require grease lubrication in close proximity, a centralized grease lubrication system can be used. A centralized lubrication system uses a single reservoir, a pump, lines to the lubrication points, and metering devices to meter grease to each bearing.




Oil Lubricated

The majority of the pumps used in Saudi Aramco installations use oil lubrication. Pumps use a variety of different methods for oil lubrication of bearings. The most common method uses an unpressurized oil system with a constant-level oiler. A constant-level oiler operates on the basis of a quantity of reserve oil stored in a bottle above the operating level of the oil in the bearing housing. The oil in the bottle is prevented from coming out since the opening of the bottle is below the operating oil level, and as long as air cannot get into the bottle, oil cannot come out. However, when the operating level drops, the opening to the bottle is uncovered, and air is allowed into the bottle. Reserve oil is thereby released until the operating level rises to cover the opening again, maintaining a constant oil level in the bearing housing. Figure 69 shows a constant-level oiler.

Figure 69. Constant-Level Oiler




Oil slingers or rings are commonly used to splash or drip oil to the bearings. The oil level in an unpressurized bearing housing should be maintained at the center of the lowermost ball of a stationary bearing. Overfilling the bearing housing with oil will result in oil leakage through the bearing seals and possible bearing overheating.

Pressurized lubrication systems are also used on pumps. Figure 70 shows a typical forced feed lubrication system used for Saudi Aramco applications. Saudi Aramco Standard 31-SAMSS-004 requires that for between bearings pumps in critical, unspared service, an external pressure lubrication systems must be in accordance with 32-SAMSS-013 and Standard Drawing AB-036858 (shown in Figure 70). For pumps that are spared, pressurized lubrication systems must be in accordance with API 610 with the following exceptions:

• The main pump must be electric motor driven.

• A separate shell and tube heat exchanger must be provided.

• The reservoir must be austenitic stainless steel.

The bearing lubrication system shown in Figure 70 consists of two lube oil pumps, a primary pump and a standby pump, that take suction from the reservoir through a strainer. Lube oil is supplied to the system through a temperature control valve (TCV). The temperature control valve maintains lube oil temperature by changing lube oil flow to the oil cooler (throttles bypass flow). The oil cooler can be either water-or air-cooled. The lube oil then passes through a duplex filter unit. If the oil cooler is water cooled, a pressure regulating valve maintains lube oil supply pressure higher than the water pressure to prevent water from leaking into the oil system from the cooler. A second pressure control valve maintains lube oil pressure to the bearings. When necessary, an overhead tank may be installed in the system to temporarily supply lube oil to the pump bearings in the event of system pressure loss until the pump coasts to a stop.




Figure 70. Typical Forced Feed Lubrication System




PUMP COUPLINGS A coupling is used whenever there is a need to connect the shafts of a prime mover to a piece of driven machinery. The function of a coupling is to axially transmit shaft-to-shaft power from one component of a machinery train to another component of the machinery train. Depending on the coupling design, a coupling may also accommodate misalignment between shafts and transmit axial thrust loads from one machine to another.

In accordance with 31-SAMSS-004, lubricated couplings are not acceptable. Lubricated couplings, such as the gear or spring grid coupling, require periodic lubrication to prevent wear of coupling parts. Nonlubricated couplings do not require grease or oil lubricant.

Flexible couplings must be capable of accepting parallel misalignment of not less than 50 mils (1270 micrometers) between shaft ends, and they must be capable of absorbing twice the maximum axial growth plus twice the axial magnetic centering force of the electric motor.

Rigid Adjustable Spacer Type In accordance with 31-SAMSS-004, rigid adjustable-type couplings are required for vertical lineshaft pumps not equipped with thrust bearings. Spacer type couplings must be furnished for all horizontal, flexibly coupled, vertical in-line and vertical suspended pumps. Couplings for high-speed integral gear pumps may be of the nonspacer type. Figure 71 shows an example of a rigid adjustable-type coupling.




Figure 71. Rigid Adjustable Coupling




Flexible Disk Pack The flexible metal disc coupling, which is shown in Figure 72, uses two hubs that are keyed to the shafts with a series of thin steel discs that are connected between the hubs. The hubs may be keyed to the shafts or hydraulically pressed onto the shafts. The flexing element is the series of thin steel discs. The series of thin steel discs is called a disc-pack. Disc-pack couplings can be configured as either single or double disc pack. A double disc-pack coupling uses two packs of thin steel discs that are attached to the hubs through a center connector ring. Each disc-pack is alternately bolted to a coupling hub and the center connector ring. High strength connecting bolts are inserted through specially designed spacer inserts that are securely held in place with lock nuts. Removal or installation of the disc-packs can be performed without removing the coupling hubs or disturbing the machinery. Disc-pack couplings do not require lubrication.

If the initial alignment of the shafts is not accurate, the assembly of the disc-packs to the coupling hubs is difficult because the closely fitted boltholes offer little clearance for bolt placement. Excessive flexing that is caused by misalignment will eventually cause the metal discs to fatigue and break.

Disc-pack couplings have the following characteristics:

• They provide high torsional rigidity.

• They can accommodate some axial shaft movement (end float).

• They can accommodate minor shaft misalignment.

• They can transmit power without any backlash or mechanical looseness between the coupling halves.




Figure 72. Double Disc-Pack Coupling

Diaphragm In accordance with 31-SAMSS-004, flexible disc-pack or diaphragm-type couplings must be supplied for all horizontal pumps and flexibly coupled vertical in-line pumps except those driven by reciprocating engines or synchronous electric motors, and vertical suspended pumps having thrust bearings.




The diaphragm-type coupling, which is shown in Figure 73, consists of two assemblies that are bolted together with a spacer. Each assembly consists of a rigid hub, a diaphragm pack, and a splined adapter. The rigid hubs are either keyed, tapered, or splined to the shafts. The outside diameter (OD) of the diaphragm pack is bolted between the hub and a diaphragm pack guard. The inside diameter (ID) of the diaphragm pack is splined to accept the splines of the splined adapter. The spline transmits the torque from the driving unit through the coupling to the driven unit. The splined adapter is held in place with a clamp ring. The clamp ring is bolted to the splined adapter with bolts that pass through the diaphragm pack. Both of the splined adapters are bolted to a spacer to complete the coupling assembly.

The flexing element is the diaphragm pack. The diaphragm pack consists of several thin convoluted diaphragms. The convolution and its unrolling action result in the diaphragm's large axial capacity with low stresses. Disc and contour diaphragms accommodate for axial motion by pure deformation of material. Removal or installation of the adapter and diaphragm packs can be performed without removing the coupling hubs or disturbing the machinery.




The diaphragm pack coupling compensates for misalignment. The diaphragm pack makes the coupling axially soft so that the coupling does not impose excessive axial loads on the bearings or the shafts of the connected equipment. The diaphragm coupling is maintenance free and it requires no lubrication. Shims may be used between the diaphragm pack and the rigid hub to adjust the effective coupling length.

Diaphragm couplings have the following characteristics:

• High and low ambient temperature capability.

• High-speed capability.

• High-torque capability.

• High-misalignment capability.

• Corrosion-resistant flex elements.

Figure 73. Diaphragm Coupling




Elastomeric In accordance with 31-SAMSS-004, elastomeric couplings or couplings that have a comparable torsional stiffness are required for pumps driven by reciprocating engines and may be considered for pumps driven by synchronous electric motors.

Elastomeric couplings consist of two hubs and a flexible elastomeric element. The hubs are fitted onto the shaft, and the elastomeric element is attached to the hubs. Figure 74 shows two of the many different styles of elastomeric couplings that are available: a splined rubber-type coupling, which is shown in Figure 74(A), and a rubber tire-type coupling, which is shown in Figure 74(B). Each style uses a different method for the attachment of the element to the hubs.

The metal hubs of the splined rubber-type coupling are either cast or are machined with a cavity that contains internal and external splines. The rubber insert also has internal and external splines. These splines engage with the hub splines. The rubber insert is a one-piece construction on small diameter couplings. On larger coupling sizes, the insert is a two-piece construction with the insert split horizontally. A steel lock ring fits into a groove, which is located on the outside diameter of the insert, to hold the insert halves together.

The rubber tire-type coupling uses two metal hubs that have grooves machined into the hubs. A flexible rubber tire element fits into the grooves and is held in place with threaded fasteners and flanges. Some rubber tire elements are molded directly to the flanges. The flanges are threaded to receive bolts. The flexible rubber element is molded into one piece, and it is then split at one point to facilitate installation.




The rubber tire coupling can be assembled and disassembled without the removal of the hubs from the equipment. The side flange bolts are removed to allow the rubber tire element to be pulled out of the flange and to allow a new rubber element to be installed.

• The elastomeric couplings have the following characteristics:

• They provide high torsional flexibility.

• They can tolerate shock loads and reversing situations.

• They act as a vibration dampener.

• They can accommodate axial shaft movements (end float).

• They can accommodate angular shaft misalignment.

• They offer excellent resistance to electrical conductivity.

• They can be used in corrosive atmospheres (provided that there is careful material selection).

Elastomeric couplings are temperature-limited, they generally have a high overhung weight, and they are difficult to balance. As the elastomers age, the assumed stiffness and damping values of the elastomer can change and reduce the flexibility of the coupling. Such reduction in the flexibility of the coupling will allow the train components to become overstressed.




Figure 74. Elastomeric Couplings




GLOSSARY

balanced seal A mechanical seal arrangement in which the effect of the hydraulic pressure in the seal chamber, on the seal face closing forces, has been modified through seal design to have a seal balance ration of less than 1.

barrier fluid A fluid that is introduced between dual mechanical seals to completely isolate the pump process liquid from the environment. Pressure of the barrier fluid is always higher than the process pressure being sealed.

bellows seal A type of mechanical seal that uses a flexible bellows to allow a static secondary seal and to provide spring-type loading on the primary ring.

buffer fluid A fluid used as a lubricant or buffer between dual mechanical seals. The fluid is always at a pressure lower than the pump process pressure.

cartridge seal A completely self-contained mechanical seal unit that contains the seal, gland, sleeve, and mating ring. A cartridge seal is pre-assembled and preset before installation.

dual mechanical seal A seal arrangement that uses more than one seal in the same seal chamber in any orientation which can utilize either a pressurized barrier fluid or nonpressurized buffer fluid.

flashing hydrocarbon service

Any service that requires vapor suppression by cooling or pressurization to prevent flashing. This category includes all hydrocarbon services where the fluid has a vapor pressure greater than 14.5 psia (1 bar) at pumping temperature.

flexible coupling A coupling that permits minor amounts of flexibility by allowing the coupling components to slide or move relative to each other.

flush A small amount of fluid that is introduced into the seal chamber on the process fluid side in close proximity to the sealing faces and usually used for cooling and lubricating the seal faces.

internal circulating device A device located in the seal chamber to circulate seal chamber fluid through a cooler or barrier/buffer fluid reservoir. Also referred to as a pumping ring.

mating ring A disk- or ring-shaped member that is mounted either on a shaft sleeve or in a housing and that provides the




primary seal when in proximity to the face of an axially adjustable face seal assembly.

neutral fluid A fluid that does not react with the process fluid. Water and steam are typically used as neutral fluids.

nonflashing hydrocarbon service

A category that includes all hydrocarbon services that are predominately all hydrogen and carbon atoms; however, other non-hydrocarbon constituents may be entrained in the stream. A product in this category does not require vapor suppression to prevent transformation from a liquid phase to a vapor phase. The vapor pressure for non-flashing hydrocarbon service is less than 14.5 psia (1 bar).

nonhydrocarbon service A category that includes all services that cannot be defined as containing all hydrogen and carbon molecules; however, some hydrocarbons may be entrained in the fluids.

non-pusher type seal A mechanical seal in which the secondary static seal is fixed to the shaft.

packing A seal that prevents leakage around the plunger or rod of a reciprocating pump.

pusher-type seal A mechanical seal in which the secondary seal is mechanically pushed along the shaft or sleeve (dynamic secondary seal) to compensate for face wear.

quench A neutral fluid (usually water or steam) that is introduced on the atmospheric side of a mechanical seal to retard the formation of solids that may interfere with seal movement.

rigid coupling A coupling that is designed to draw the two shafts together tightly so that no relative motion can occur between the shafts.

seal balance ratio Sometimes expressed as a percentage, the ratio of seal face area that is exposed to closing force by hydraulic pressure in the seal chamber, to the total sealing face area.

seal chamber A component that is either integral with or separate from the pump case (housing) and that forms, between the shaft and casing, the region into which the shaft seal or packing is installed.

specific gravity The density of a liquid divided by the density of water at 60°F.

Seal Systems, Bearing Arrangements and Couplings in Saudi A

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