Volume 18 - Friction, Lubrication, and Wear Technology Part 16 ppt

High-energy braking conditions produce a high coverage of wear debris film, which may result in a low coefficient of friction and low wear rate.. Table 1 Commonly used abradable seal mat

Fig 6 Variation in wear of carbon brake from taxi to RTO energies

Each of the three energy ranges has its own wear mechanism The taxi region exhibits the highest wear, and mechanical abrasion predominates In the mid-energy (service energy) region, wear is lowest because a thin film protects the substrate The third region occurs when temperatures are sufficiently high to produce oxidation of any debris or substrate (fiber or matrix) Oxidation usually occurs during overload and RTO conditions

PAN-base fibers wear differently than pitch-base fibers; these differences are associated with fiber processing as well as composite processing A harder matrix will wear less than a softer one Studies have shown that circumferentially oriented fibers have high wear and a high coefficient of friction

Wear Mechanism and Wear Debris Analysis. The worn surfaces of the carbon materials have bands of different reflectance when observed via macro examination Debris material occurs in both bright and dull bands The observed banding on the wear surface of the composite is related to variations in reflectance; these variations are thought to be related to differences in the character of the wear debris The bright bands consist of a thin film of debris with a polished appearance that produces a high reflectance The dull bands have a high number of fissures or shallow grooves and scratches in the longitudinal fiber bundles, both of which are indicative of fiber removal caused by abrasion The debris material in the dull bands does not produce a high reflectance and appears less dense (more porous) than the bright bands The debris in the dull bands is particulate and is comprised of fibers and matrix This debris does not form a film, but fills

in the original porosity, mainly in the matrix, and is found primarily under low-energy conditions

The debris in the bright band consists of a thin film (up to a few micrometers thick) that is smeared over fibers and matrix and also fills in the pores The bright bands appear to be denser (less porous) than the dull bands The wear debris film is grooved, whereas the fibers and matrix are not This indicates that the wear debris film is protecting the fibers and matrix from abrasion

The thin film of wear debris is amorphous in character and is comprised of both fibers and matrix High-energy braking conditions produce a high coverage of wear debris (film), which may result in a low coefficient of friction and low wear rate More detailed information on wear debris analysis can be found in the article "Lubricant Analysis" in this Volume

Moisture Problems. Carbons and graphites have an affinity for moisture These materials adsorb moisture; that is, water molecules are attracted to their exterior and interior surfaces Consequently, the presence of moisture significantly reduces the coefficient of friction When a brake is sitting for a period of several hours, the rubbing surfaces will adsorb moisture from the air Thus, when a brake stop is initiated in this condition, the stopping power is significantly reduced because of the low coefficient This is typically referred to as "morning sickness." As rubbing continues through the stop, the moisture evaporates and the coefficient returns to its normal dry value

Oxidation. Carbons and graphites are also subject to oxidation at elevated temperatures Typically, the threshold of

oxidation is considered to be 430 °C (805 °F) Technically there is oxidation at this temperature, but it is so low that it is considered to be negligible Other than taxi conditions, the operating temperatures of a carbon brake will range between

500 and 1000 °C (930 and 1830 °F) during landings; under RTO energies the temperatures exceed 1300 °C (2370 °F) and oxidation becomes a significant factor In order to keep the disks from deteriorating, an oxidation inhibitor is applied to the outer and inner diameters of each carbon brake disk Inhibitor is not applied to the wear surfaces because it would alter the friction characteristics of the carbon

Friction Coefficient Variability. As stated previously, friction coefficients have a larger range over the operational spectrum of the aircraft when compared to the friction coefficients for steel brakes This coefficient range has implications for brake control (antiskid) system design In the carbon friction material, friction coefficients can vary by factors of 3 or more over the operation range of the brake Therefore, the brake torque can vary by a factor of 3 or more Carbon processing and brake frame design must take this friction coefficient variation into account

Vibration. An aircraft wheel and brake is a system with multiple degrees of freedom that is subjected to high dynamic loads This high dynamic loading is transient in nature and can be an exciter of vibration

In a carbon brake, the most critical vibration mode is known as whirl This mode consists of an accordian-type action of the brake disks combined with an orbiting motion of the brake structure about the axle (Fig 7) The torque output and friction coefficient of the brake, as functions of velocity and brake pressure, are significant parameters in determining whether or not whirl motion will occur and, if it does occur, the severity of the vibration Severe whirl can result in damage to the carbon brake disks and other wheel/brake hardware

Fig 7 Carbon brake vibration

There are two basic approaches that can be taken to control whirl vibration First, carbon friction coefficient characteristics are engineered in such a way that the brake coefficient, at a given velocity and brake pressure, remains low enough to minimize the potential for the whirl vibration Second, the aircraft wheel/brake is designed to incorporate the required stiffnesses and damping characteristics Adequate stiffnesses and damping, combined with favorable friction coefficient characteristics, will ensure the stability of the aircraft wheel/brake structure at given dynamic conditions

Testing

Friction material testing is conducted on either direct-connected dynamometers or on landing wheel dynamometers Typically, new friction materials are screened in subscale brakes on direct-connected dynamometers Once a potential material has been selected, then full-scale brakes are tested on landing wheel dynamometers which use an aircraft tire and wheel Full-scale brake testing is very expensive and therefore limited During aircraft wheel and brake qualification testing, the full-scale brake is run through numerous tests before any aircraft testing is done

Test Requirements. Generally an aircraft brake must pass testing standards set up by the military or FAA The military requirements are outlined in MIL-W-5013, and the FAA requirements are continued in TSO-C26 The Society of Automotive Engineers (SAE) also publishes an Aerospace Recommended Practice (ARP) for aircraft wheels and brakes (ARP 597) The airframe manufacturers also specify extensive supplemental qualification requirements that must be met before the brake can be qualified for service

Selected References

• J.F Archard, The Temperature of Rubbing Surfaces, Wear, Vol 2, 1958/59

• F.P Bowden and J.E Young, Proc R Soc (London) A, Vol 208, 1951, p 444

• H.W Chang, Correlation of Wear with Oxidation of Carbon-Carbon Composites, International

Conference on Wear of Materials, American Society of Mechanical Engineers, 30 Mar to 1 Apr 1981

• T.S Eyre and F Wilson, Wear of Grey Cast Iron Under Unlubricated Sliding Conditions,

ASME/ASLE International Lubrication Conference (New York), 9-12 Oct 1972

• D.B Fischbach and D.R Uptegrove, Oxidation Behavior of Some Carbon/Carbon Composites, 13th

Biennial Conference on Carbon (Irvine, CA), 1977

• K Gopinath, G.V.N Rayudu, and R.G Narayanamurthi, Friction and Wear of Sintered Iron, Wear,

Vol 42, 1977, p 245-250

• B Granoff, H.O Pierson, and D.M Schuster, Carbon-Felt, Carbon-Matrix Composites: Dependence

of Thermal and Mechanical Properties on Fiber Volume Percent, J Compos Mater., Vol 7, Jan 1973

• T.-L Ho, "Development and Evaluation of High-Energy Brake Materials," Ph.D thesis, Rensselaer

Polytechnic Institute, 1974

• J.M Hutcheon and M.S.T Price, The Dependence of the Properties of Graphite on Porosity,

Proceedings of the Fourth Conference on Carbon, Pergamon Press, 1960

• W.V Kotlensky and P.L Walker, Jr., Crystallographic and Physical Changes of Some Carbons Upon

Oxidation and Heat Treatment, Proceedings of the Fourth Conference on Carbon, Pergamon Press,

1960

• I.V Kragelskii, Friction and Wear, Butterworths, Washington, 1965, p 117

• J.K Lancaster, Instabilities in the Frictional Behavior of Carbons and Graphites, Wear, Vol 34, 1975

• R.L Lewis and R.E Raymond, "Stopping Distance Analysis," Society of Automotive Engineers, Inc.,

Paper No 730193, 1973

• F.F Ling and E Saibel, On Kinetic Friction Between Unlubricated Metallic Surfaces, Wear, Vol 1,

1957/58

• J Molgaard and V.K Srivastava, The Activation Energy of Oxidation in Wear, Wear, Vol 41, 1977

• N Murdie, C.P Ju, J Don, and F.A Fortunato, Microstructure of Worn Pitch/Resin/CVI C-C

Composites, Carbon, Vol 29, 1991, p 335-342

• D Pavelescu and M Musat, Some Relations for Determining the Wear of Composite Brake Materials,

Wear, Vol 27, 1974

• T.F.J Quinn, A.R Baig, C.A Hogarth, and H Muller, Transitions in the Friction Coefficients, the

Wear Rates, and the Compositions of the Wear Debris Produced in the Unlubricated Sliding of

Chromium Steels, ASME/ASLE International Lubrication Conference (New York), 9-12 Oct 1972

• E, Rabinowicz, Friction and Wear of Materials, John Wiley & Sons, 1965

• D.M Rowson, The Interfacial Surface Temperature of a Disk Brake, Wear, Vol 47, 1978

• L Rozeanu, Friction Transients (Their Role in Friction Failures), Trans ASLE, Vol 16, 1975, p

257-266

• J.J Santini and F.E Kennedy, Jr., An Experimental Investigation of Surface Temperatures and Wear

in Disk Brakes, Lubr Eng., Aug 1975

• P Stanek, N Murdie, E.J Hippo, and B Howdyshell, The Effect of Fiber Orientation on Friction and

Wear of C-C Composites (Extended Abstracts), Biannual Conference on Carbon (Santa Barbara, CA),

1991, p 378-379

• I.L Stimson and R Fisher, Design and Engineering of Carbon Brakes, Philos Trans R Soc

(London) A, Vol 294, 1980

• E.M Tatarzycki, Friction Characteristics of Some Graphites and Carbon Composites Sliding Against

Themselves, 13th Biennial Conference on Carbon, 1977

• A.K Vijh, The Influence of Solid State Cohesion of Metals and Non-Metals on the Magnitude of

Their Abrasive Wear Resistance, Wear, Vol 35, 1975

Wear of Jet Engine Components

J.D Schell and K.P Taylor, General Electric Aircraft Engines

Introduction

A JET ENGINE is a sophisticated piece of machinery with many moving parts; the potential for wear problems exists whenever moving parts come into contact or unintended motion occurs between stationary contacting parts As for any system, wear in jet engines can be controlled through proper design, material selection, and lubrication A schematic cross section of a typical jet engine is shown in Fig 1 The major engine subsystems consist of the fan, the high-pressure compressor (HPC), the combustor, the high- and low-pressure turbines (HPT and LPT), and the exhaust nozzle The engine design contains one nonrotating system and two concentric rotating systems The nonrotating (stator) system is made up of structural frames and casings The low-pressure rotating system consists of the fan disk(s) and fan blades, the LPT disks and turbine blades, and a connecting shaft The high-pressure rotating system consists of the HPC disks/spools and compressor blades, the HPT disks and turbine blades, and a connecting shaft

Fig 1 Jet engine cross section showing important subsystems and potential areas of wear

Operating environments vary widely between different sections of the engine and depend on where the engine is in its mission Temperatures may vary from subzero to above 1095 °C (2000 °F), rotational speeds may climb to more than 15,000 rev/min, and contact loads may range from a few psi to local hertzian stresses well beyond 1720 MPa (250,000 psi) in rolling-element bearings The relative motions of components may be unidirectional sliding of rotating parts on stators, oscillatory sliding varying from a few thousandths of an inch up to several tenths of an inch, or vibratory motion resulting in impact between components Components also may be subjected to ingested particle impacts

The wide variety of operating conditions results in a wide variety of materials used to meet the design needs of the engine Aluminum and titanium alloys, plastics, and resin-matrix graphite composites are frequently used in the fan and the engine nacelle The HPC uses titanium alloys, nickel-base superalloys, such as Inconel 718, and steels, such as M152, 17-4PH, and A286 The combustor requires heat-resistant nickel or cobalt alloys, such as Hastelloy X or Haynes 188, and stainless steels for fuel tubing The turbine sections rely on cobalt and nickel superalloys, such as Inconel X750, MAR-M-

509, René 77, René 80, René 125, and advanced directionally solidified and single-crystal alloys Often the design demands on materials for jet engine components will not permit substitution of materials for wear purposes, so a number

of surface coatings and treatments are employed for wear protection

The different operating environments and types of materials in each section of the engine result in a variety of wear types, including fretting, impact, adhesive, high-speed and oscillatory sliding, oxidational, ingested particle erosion, and abrasive wear High-speed sliding wear occurs in rotating gas path seals Impact wear can occur in loose part assemblies

or blade midspan or tip shroud interlocks Fretting wear is frequently seen in blade dovetails Erosion occurs when dirt and sand particles are ingested with the air through the fan and compressor Bearings and gears can experience rolling contact fatigue High-temperature components can experience oxidational wear This article will discuss some of the most significant of these wear problems in relation to specific jet engine components

Gas Path Seals

A major area of wear in jet engines involves gas path sealing Such seals include blade tip seals, labyrinth seals, and leaf and spline seals Blade tip and labyrinth seal problems are concerned with clearances between rotating parts and their adjacent stators Engine efficiency is significantly affected by the amount of gas leakage over blade tips or through labyrinth seals In an ideal engine, the blade tips or labyrinth seal teeth would maintain minimum clearance with the adjacent stator surfaces at all points in the engine cycle In practice, the rotor parts and stator parts experience differential growth rates because of thermal gradients in the engine and much larger mechanical growth of the rotor than the stator because of centrifugal forces The results of these differences are depicted in Fig 2 When the rotor and stator diameters are plotted against time after the throttle is applied for takeoff, they are seen to experience a period of interference (pinch point in Fig 2) that causes wear The issue can be further complicated by the stator going out of round (Fig 3)

Fig 2 Rotor and stator growth rates as a function of time and engine throttle movements

Fig 3 Clearance change caused by rotor/stator eccentricity or maneuver deflections

Both design and material approaches are employed in combating wear of blade tips One design approach, known as active clearance control, applies heating or cooling to the stator to achieve a better match of the thermomechanical responses of the stator and rotor A second design approach for the out-of-round condition involves local arc grinding to remove casing material in the areas where minimum radii would occur A materials approach that has been used in recent years is called passive clearance control The casing is made of an alloy with a low coefficient of expansion, such as IN909, to achieve a more favorable overall thermal transient response match of the rotor and stator diameters The most commonly used materials approach involves the application of an abradable material to the stator The abradable material wears preferentially in a limited arc when the stator is out of round or when the rotor moves off center This results in local clearance increases during rotor/stator interferences instead of wearing the rotor and causing a 360° clearance increase An alternative materials approach is to apply an abrasive to the rotor, which machines the stator material, thus achieving the same result

The materials used for the abradable stator seals or abrasive rotor coatings vary by location in the engine Abradables can take several forms, including bonded elastomers, braze-attached sintered porous metals or honeycomb cells, or thermal spray coatings Some of the more commonly used abradables are listed in Table 1 These materials are designed to wear

in preference to the opposing blade tip or seal tooth They rely on low densities created by included porosity or friable structures with weak bonding between constituent materials Bill and Wisander (Ref 1) have provided a model for friable abradable seal materials In practice, however, wear usually occurs on both surfaces, necessitating periodic overhaul

Table 1 Commonly used abradable seal materials

Type of seal Material

Phenolic/carbon microballons Aluminum

80/20 nickel-graphite Porous Teflon Aluminum-silicon/polyester

Fan and booster seals

Ni-Cr-Al/bentonite Nickel-graphites (75/25, 80/20, and 85/15) Nickel-aluminum

Aluminum Aluminum bronze/nickel-graphite Ni-Cr-Al/nickel-graphite

Ni-Cr-Al/bentonite Hastelloy X open-faced honeycomb

High-pressure compressor seals

FiberMetal Co-Ni-Cr-Al-Y

High-pressure turbine seals

Bradelloy (Hastelloy X honeycomb + braze/nickel-aluminum

The abrasive materials approach has been used with success on rotating parts, allowing them to machine their own clearances and minimizing rotor wear The most commonly used abrasive is plasma spray aluminum oxide on seal teeth

or rotor lands Figure 4 shows a sector from an HPC rotor with two sets of seal teeth coated with plasma spray aluminum oxide The most common mating stator seal material for such applications is open honeycomb (Fig 5) Commonly used abrasive coatings for clearance control in jet engines include:

• Plasma spray aluminum oxide

• Entrapment-plated cubic boron nitride (Borazon)

• Entrapment-plated aluminum oxide

The abrasive coatings approach is usually combined with honeycomb or an abradable seal to improve the overall wear system for both surfaces

Fig 4 High-pressure compressor disk with seal teeth coated with plasma spray aluminum oxide between

stages Arrow indicates location of coated seal teeth

Fig 5 Open-faced honeycomb seal showing cutting by seal teeth

Blade Midspan Stiffeners and Tip Shrouds

Some fan, HPC, and LPT rotating airfoils (blades) require the use of either a midspan stiffener or a Z-notch tip shroud (often called interlocks) to prevent mechanical flutter of the aerodynamically loaded blades These must be designed so that the blades are sufficiently loose to allow easy assembly, but lock up into a solid stiffening ring as aerodynamic loads are imposed on the blades, causing them to untwist along the blade stacking axis These two requirements result in a combination of impact and sliding as the interlocking contact surfaces engage and rotate into position to form the solid stiffening ring

The impact loads imposed on the contact surfaces can be on the order of 7 to 70 MPa (1000 to 10,000 psi) and can cause severe wear damage to most materials suitable for use as blades Therefore, it is common practice to apply a wear material to the interlock contact surfaces These wear treatments are usually coatings on the order of 0.13 to 0.25 mm (0.005 to 0.010 in.) thick or welded hardfacing deposits up to 2.5 mm (0.100 in.) thick Much care must be taken in the design and assembly of alignment tolerances for interlocks to prevent excessive wear, chipping, or spallation of even the most successful wear treatments on the interlock contact surfaces

The materials used for fan and HPC blades with interlocks are usually titanium alloys, which have poor wear properties Most fan blade and HPC interlocks use thermal sprayed WC-Co coatings or brazed-on WC-Co powder metallurgy wear pads to prevent excessive wear The most widely used coating is Union Carbide's LW1N40, applied using a detonation gun (D-gun) Recent advances in thermal spray coatings have allowed the use of high-energy plasma spray WC-Co coatings, which hold promise for direct substitution, or high-velocity oxyfuel (HVOF) sprayed WC-Co coatings on titanium alloy interlocks The WC-Co coatings are successful in the titanium alloy interlock applications because of the high wear resistance of the tungsten carbide, adequate fracture toughness because of the cobalt matrix, high adherence on the titanium alloy substrates, and a good match in coefficient of thermal expansion with the titanium alloy substrate materials The typical range of temperatures for fan and HPC interlocks may vary from subzero to 95 °C (200 °F) in the fan and from about 40 to 260 °C (100 to 500 °F) in the HPC Fortunately, WC-Co coatings appear to retain sufficient low-temperature ductility and high-temperature oxidation resistance over these temperature ranges The formation of a wear glaze at the contact zones contributes to the good wear resistance of the WC-Co in these interlock applications

The LPT blade materials are typically nickel-base superalloys, such as René 77 or René 125, which usually possess fairly good sliding wear resistance However, they have inadequate wear resistance in the combined impact and sliding wear environment of LPT blade interlock contact surfaces Typical use temperatures for LPT interlocks are 540 to 925 °C (1000 to 1700 °F), so the oxidation properties of the alloys under the existing wear conditions also play a significant role

in their wear resistance Typical wear coating compositions applied by thermal spraying or weld buildup that are used for LPT blade interlocks include:

• Tribaloy 800 (plasma sprayed, welded, HVOF)

• Cost Metal 64 (welded)

• Chromium-carbide/nickel-chromium (plasma sprayed, HVOF, D-gun)

Most of these alloys are cobalt based for good wear resistance and benefit from the formation of cobalt oxide and/or spinel wear glaze films

Tribaloy 800 and Coast Metal 64 are the most commonly used LPT blade interlock coatings at GE Aircraft Engines Tribaloy 800, applied by thermal spraying or tungsten inert gas (TIG) welding, provides excellent wear resistance and oxidation resistance to about 840 °C (1550 °F) Above this temperature TIG-welded Coast Metal 64 provides better wear and oxidation resistance than Tribaloy 800 In general, wear coating performance for LPT blade interlocks correlates to the chromium content and use temperature, with better performance at elevated temperatures for coatings with higher chromium contents and better performance at lower temperatures for coatings with lower chromium contents

Mainshaft Bearings

The materials traditionally used for gas turbine mainshaft bearings are 52100 and M50 steels More recently, powder metallurgy (P/M) bearings and case-carburized M50NiL steel, a modified M50, have been introduced as race materials Several factors have contributed to this recent trend Newer gas turbine mainshafts operate at higher speeds This has pushed the bearing DN values (bore diameter in millimeters times shaft revolutions per minute) well past 2 million, which increases race hoop stresses and the hertzian contact stresses between rolling elements and the races

The higher hoop stresses can cause fracture of the 52100 or M50-type races because they are through-hardened materials, typically in the 50 to 60 HRC range, and thus have low fracture toughness At high DN values, this can become a fracture reliability problem for a statistically significant number of these bearings The higher hertzian stresses, approaching 2400 MPa (350,000 psi) for a 2.5 million DN mainshaft bearing outer race, also can cause significant reductions in rolling contact fatigue life This is undesirable, because changing a mainshaft bearing requires costly disassembly of the engine

Bearing races made from the new P/M alloys and forged low-carbon alloys with carburized surfaces do not have these shortcomings These materials are designed for high DN use and require special manufacturing processes GE Aircraft Engines has concentrated on a variation of M50 steel with reduced carbon and increased nickel to improve fracture toughness The race is then carburized to produce a fine dispersion of carbides for high hardness Compressive residual stresses are frozen in to the raceway surfaces to improve rolling contact fatigue, while the low hardness (<50 HRC) rare core material remains tough to deal with high hoop stresses

This M50NiL material with its finely dispersed carbides, as well as the fine-grained (to improve fracture toughness) P/M race materials, can suffer from a low tolerance to wear Wear can occur at the ball cage guide lands under marginal lubrication conditions even for "normal" bearing cleanliness operation The rolling-element cage shoulders are silver plated, which provides solid lubrication and low friction to prevent wear when direct metal-to-metal cage skidding occurs

on the cage guide land of the race This works quite effectively for 52100 and M50 steels However, the P/M and M50NiL steels sometimes experience rapid wear under similar operating parameters

Research by Budinski (Ref 2) has shown that the size, distribution, type, and volume fraction of carbides in tool steels can significantly alter their abrasive wear resistance, with coarser carbide grains having better resistance than finer carbide grains Thus, it has been suggested that the coarser carbide stringers in M50 or 52100 forged bearing races can adequately resist the initiation of abrasive wear, while the very fine; evenly dispersed carbides in M50NiL or P/M bearing races cannot

The abrasive particles found in the bearings originate in the oil supply system Sump castings, abrasively cut tubes, and grit-blast-cleaned parts in the bearing lubrication system all likely contain very fine alumina or silicon carbide contaminant particles Most of these are removed during cleaning prior to assembly or by in-line filtering, but some of the finer particle (<50 m) get through to the bearings even under the most stringent clean-room assembly conditions Once inside a bearing, contaminant particles can become embedded in the silver plating on the ball cage shoulders, where they protrude, causing abrasive wear to initiate during transient cage shoulder/race guide land contacts The abrasive particles soon become "capped" with race transfer material, and adhesive wear ensues

The high differential sliding speeds between the orbiting cage and race cause frictional heating, local oxidation, and carburization by oil coking of the thin metallic transfer layers Thus, a very hard abrasive transfer layer results and the wear process accelerates These deposits also increase friction Therefore, when cage-to-cage encounters occur, a more severe rebound force results from the skidding contact, generally increasing the cage orbiting and number of skid contacts and producing further wear

Combustor and Nozzle Assemblies

These engine components are subjected to a variety of severe wear environments Combustor hardware includes fuel nozzles, swirlers, and cowl damping wires, which experience relatively high temperatures (540 °C, or 1000 °F, and up) during operation Exhaust nozzle assemblies are characterized by many parts, such as pins, bushings, links, and overlapping flaps, which aid in motion of the nozzle to control engine thrust Some of these nozzle parts are directly in the hot gas stream and experience temperatures up to 815 °C (1500 °F); others are bathed in bypass cooling air and remain relatively cool (approximately 315 °C, or 600 °F) Combustor and nozzle assemblies experience large amounts of vibration from turbulent air flows both inside and outside the engine This vibrational/impact wear can cause significant material removal as well as high-cycle fatigue of some components The combined effects can cause liberation of hardware; in the case of the combustor, this will in turn cause severe damage to downstream components, such as turbine nozzles and blades

The majority of wear problems in both of these assemblies is cause by vibration and impact Because of the elevated temperatures in the combustor, oxidational wear occurs as scales are formed and subsequent chipped off by impact Contact pressures between parts are nominally low, but can be aggravated by high-frequency impacts, which may locally yield the materials Oscillatory sliding (galling) wear sometimes occurs on exhaust nozzle flaps as they are actuated during mission cycles over several hundred accumulated flight hours

The design of these components addresses temperature and fatigue concerns Both combustor and exhaust nozzle hardware are made from heat-resistant superalloy sheet materials, such as Hastelloy X, René 41, or Haynes 25 Because these materials vibrate in the turbulent hot gas stream, high-cycle fatigue life at elevated temperature is important The cooler sections of the exhaust nozzle sometimes use high-temperature titanium alloys, such as Ti-6Al-2Sn-4Zr-2Mo, to reduce engine weight and maintain mechanical properties at elevated temperature Many pins and bushings are manufactured from steel alloys, such as 17-4PH and A286 Coatings can be applied to problem areas on specific components, but they must withstand the application temperatures and not degrade the mechanical properties of the base alloy to unacceptable levels Therefore, specification of the material and/or coating can be a complicated process

In general, cobalt-base alloys, such as Haynes 25 and Haynes 188, tend to perform best at temperatures above 540 °C (1000 °F) in both sliding and impact wear Some nickel-base alloys, such as René 41, also possess good wear resistance at high temperatures Effective coatings for wear problems in these temperature regimes include Tribaloy 800 and chromium carbide/nickel chromium Tribaloy 800 derives its good elevated-temperature wear resistance from a hard Laves phase in a cobalt-base matrix This matrix produces a cobalt oxide, which provides lubricity to the interface Chromium carbide/nickel chromium derives its good performance chiefly from the hard carbide phase and the formation

of favorable oxide wear glazes Cooler titanium components in the exhaust nozzle generally have poor wear resistance and almost always require coatings for mating parts in relative motion Here, the coating of choice is generally WC-Co, which again derives its wear resistance from the hard carbide phase Oxidation of the carbide limits use of this coating to temperature regimes below 480 °C (900 °F) Because of the aggressive nature of the carbide, both mating surfaces should

be coated The steels used in the actuation systems for the nozzle flaps are usually ion nitrided to develop a hard case layer on the surface (hardness of up to 72 HRC can be achieved) For a particularly severe wear environment, ion nitriding may not provide sufficient wear protection, and chromium carbide or tungsten carbide coatings may be required

Dovetails

Aircraft engines utilize bladed rotors in the low-pressure fan and the high-pressure compressor to generate thrust and provide the required air compression ratios for proper fuel combustion to drive the turbine sections These airfoils are attached to their respective disks on the rotor by dovetails (Fig 6) As the rotor speed increases, the disk posts grow and open while the mass of these blades is forced radially outward by centrifugal forces These forces result in relative motion between the blades and the disk under very high normal stresses, reaching localized peak loads of up to 930 MPa (135 ksi) in the fan

Fig 6 Front-stage HPC rotor disk showing dovetail disk slots and pressure faces

Superimposed on some stages of the high-pressure compressor are high-frequency, low-amplitude vibrations caused by local aerodynamic effects from vane or strut positions relative to the rotating hardware The combination of contact stress and high-frequency motion admits a large amount of mechanical work into the metal surface Temperatures for dovetails range from subzero (fan) to close to 650 °C (1200 °F) (rear compressor) The combination of these factors creates wear problems for critical rotating hardware

Fan and front-stage compressor blades and disks are manufactured from titanium alloys; middle- and rear-stag compressor rotor materials included steel (A286) and nickel-base alloys (Inconel 718) Rear-stage components rarely exhibit wear; however, titanium alloys have notoriously poor wear resistance when unprotected Any cause of surface damage, such as wear, must be minimized because of the surface sensitivity of titanium in relation to high-cycle fatigue crack initiation

Most jet engines in service use a titanium protection system comprising a thermal spray coating and a solid film lubricant

on the blade dovetail Most disk and spool dovetail slots cannot accommodate a thermal spray coating in the disk slots because they are relatively small, but the largest fan disk slots use the same coating system as the blade dovetails The coating of choice is a relatively soft copper-nickel or copper-nickel-indium alloy applied by air plasma spraying The thermal spray coating is subsequently dry film lubricated with molybdenum disulfide in an organic resin binder that is oven cured for strength and adhesion The solid film lubricant provides a low coefficient of friction to prevent galling, allow part seating, and delay the onset of wear produced by metal-to-metal contact Coating thicknesses range from 0.05

to 0.28 mm (0.002 to 0.011 in.) for the thermal spray and 0.025 to 0.075 mm (0.001 to 0.003 in.) for the cured lubricant

Titanium galls readily in contact with other metals in relative motion, even against the soft thermal spray coating, so solid film lubricant life is important Unfortunately, copper-nickel and copper-nickel-indium alloys and MoS2 can oxidize above about 315 °C (600 °F), while the organic binder breaks down and loses strength at such temperatures, resulting in higher friction coefficients, increased shear forces, and wear Improved dovetail coatings for the 315 to 540 °C (600 to

1000 °F) range need to be developed, as new high-temperature titanium alloys are being specified farther back in the compressor

Compressor Airfoil Erosion

The introduction of underwing-mounted jet engines has brought with it the problem of the ingestion of materials left on airport runways Sand and dirt are used during winter months to improve traction on snow-covered runways This grit is easily ingested by the engine when the snow melts and evaporates Because the current "hub-and-spoke" airline system produces many takeoff/landing cycles in a short time, compressor blades and vanes erode rapidly from these contaminants Erosion is also prevalent in helicopter engine compressor because of the harsh unpaved environments in

which they operate Helicopter engines have particle separators to eliminate very large grit sizes, but finer grit sometimes enters the compressors This particle ingestion is severe enough to cause erosion of titanium-, nickel-, and steel-base materials to various degrees

Erosion occurs by two distinct mechanisms: high-impingement angle and low-impingement angle angle erosion is characteristic on the front stages of the high-pressure compressor when large ingested particle strike the leading edges of the rotating blades In the most severe form of this type of erosion, the leading edge can be deformed by burring or plastic deformation Otherwise, high-angle-impingement attack usually results in blunting of the leading edge Both changes in leading edge profile cause aerodynamic flow problems that reduce compressor efficiency and increase fuel consumption

High-impingement-Low-impingement-angle erosion is characteristic on the rear stages of the high-pressure compressor As the ingested material swirls through the compressor, it becomes fragmented and centrifuged radially toward the casing This results in glancing impacts by angular, sharp grit near the trailing edge of the airfoil and along the tip The loss of airfoil material causes thinning of the trailing edge and opens tip clearances near the casing, reducing compressor performance Figure 7 illustrates the mechanism of low-impingement-angle erosion

Fig 7 Rear-stage compressor blades illustrating the low-impingement-angle erosion mechanism Arrow

indicates lost erosion coating on blade trailing edge

Protection of individual blades has historically consisted of coating those airfoil surfaces most affected by erosion: the leading edge, across the concave surface, and back to the trailing edge in the tip region of the blade Based on field experience, coating practice has been modified to include wrapping the coating around the leading edge and extending the coating down the airfoil surface in the direction of the dovetail Coating materials of choice are those with higher hardness than the impinging grit, such as carbides and nitrides of titanium, tungsten, and chromium These coatings can

be applied by processes such as physical vapor deposition (PVD), D-gun, or HVOF thermal spraying to maximize their density and hardness

The chosen coating system must not affect the mechanical properties of the base alloy, such as high-cycle fatigue resistance Blade airfoils pass through several resonant frequencies during an engine cycle Degradation of fatigue strength by the coating could cause cracking and liberation of the blade tip, producing more damage downstream The coating system must also exhibit excellent adhesion to resist spallation and produce a very smooth surface finish for compressor efficiency Any coating-induced surface defect will cause unacceptable turbulence, which can disrupt the axial flow through the compressor Coating thicknesses are relatively thin, typically less than 0.025 mm (0.001 in.) for PVD coatings and 0.050 to 0.075 mm (0.002 to 0.003 in.) for D-gun and HVOF coatings Coatings must remain thin so that they taper smoothly to the uncoated blade, without a sharp step that would cause compressor airflow turbulences

Some helicopter engine compressor are constructed from one-piece blade and rotor components, called blisks These components do not allow for the coating of individual airfoil surfaces, because the blades are an integral part of the disk (Fig 8) Blisks also have twisted airfoils in the direction of the stacking axis, making them difficult to coat using a line-of-sight process such as thermal spray Leading edges are extremely thin, sharp, and prone to severe burring and blunting under particle impingement Development of non-line-of-sight erosion coatings for difficult geometries such as these is required to maintain helicopter performance

Fig 8 Front-stage helicopter engine compressor rotor blisk

Discussion and Summary

Jet engine designs push components to their limits for mechanical strength and temperature capability New engine designs for commercial airplanes and military fighter jets introduce new requirements for thrust and performance The fact is that the hotter the combustion in the engine, the more powerful and efficient the engine becomes Advances such as new superalloys, improved cooling flow designs, and thermal barrier coatings were developed to stretch the capabilities

of these components and to take advantage of the thrust and efficiency gains associated with increased combustion temperatures

A second method of generating more thrust is to increase the size of the fan and thus of the turbine required to drive it Increased size results in larger structures with more thermal movement Relative motion and vibration are also expected

to increase with these new designs

To accommodate size increases without substantially increasing weight, lighter materials are being developed with similar strengths to handle these more demanding conditions of temperature and loading These advanced materials include ceramics, organic-matrix composites, metal-matrix composites, ceramic-matrix composites, and intermetallics Each class has its own property challenges that must be overcome before it can be incorporated in engine designs For example, composites utilize second phases to improve properties lacking in the matrix material, such as tensile strength and ductility Intermetallic and ceramic materials can be very sensitive to any type of surface defect that may cause brittle fracture, including wear or erosion

These advanced mechanical designs and material systems provide research opportunities in the field of tribology The effects of increased temperature, of increased motion and vibration, and of second phases in composites on the wear process require further investigation

References

1 R.C Bill and D.W Wisander, Friction and Wear of Several Compressor Gas-Path Seal Materials, NASA

TP-1128, National Aeronautic and Space Administration, 1978

2 K.G Budinski, Surface Engineering for Wear Resistance, Prentice-Hall, 1988

• Erosion caused by high-frequency impact damage from liquid cavitation bubble formation and collapse

in the flow stream of locations where the static pressure is locally reduced below the vapor pressure

Pitting corrosion is beyond the scope of this article The rubbing wear, particulate erosion, and cavitation erosion mechanisms are discussed separately below Each discussion focuses on specific components subject to the given form of wear and includes:

• The effects of wear on overall pump performance

• The relationship of the wear mechanism to relevant design parameters and environmental variables

• Specific defenses that minimize the rate of material loss, and hence extend the life of the pump, or that affect the choice of a specific pump

The pumps discussed in this article are primarily heavy-duty industrial types, including single-and multistage centrifugal, single-stage axial flow, reciprocating, and positive-displacement progressive cavity pumps, such as screw pumps

• Interstage bushing or diffuser bushing annular seals, which separate the inner diameter of the backside

of one stage from the inlet area of the following stage

• Axial thrust-balancing devices, such as drums with a radial annular seal or disks with an axial face seal Such balancing devices separate multistage discharge pressure acting on one cross-sectional area from suction pressure acting on another cross-sectional area in order to oppose the accumulated axial thrust of the various pump stages In the drum type, which is generally preferred because of its dynamic stability,

if the original running clearance opens up sufficiently, parasitic leakage becomes unacceptable, and pressure values on each side change to the point that thrust-balance may no longer be possible

• Sealing between the casing fluid and the outside environment, such as "pressure-breakdown" bushing labyrinth seals, pump packing, and mechanical seals Such seals are eliminated in magnetic-drive pumps, canned motor pumps, and most other types of hermetically sealed pumps

• Flexible couplings involving potentially sliding teeth or splines, such as gear couplings

• Bearings, including plain journals, fixed pad, tilting pad, and rolling-element bearings

Fig 1 Cross-sectional view of a multistage pump showing wear-prone components Courtesy of Dresser Pump

Division, Dresser Industries

These components represent the closest clearance locations within a centrifugal pump and, even in an upset condition in which one or more of them is rubbing, prevent the rotor from contacting at other locations This same principle prevents excessive bending deflection of the rotor, such that shaft lateral steady deflection or lateral vibrations cannot induce sufficient bending stress to result in fatigue failure of the internal shaft (Ref 1) Of course, a serious metallurgical defect

or machining imperfection that causes a stress concentration is always a potential failure source, whether or not there is a vibration problem

It is common for a centrifugal pump to experience wear in one or more of the locations listed above The goal of the pump designer and pump user should be to minimize rather than prevent rubbing wear at close clearances, so that the original clearances are at most tripled during intervals between pump overhauls Therefore, the effects of up to three times increased clearance at all potential rubbing locations on hydraulic performance degradation must be accounted for in pump design Generally, roughly one to two points of efficiency and 0.5 to 1% of head are lost when running clearances throughout a pump are doubled

The predicted rotor dynamic behavior of the pump as clearances change is also critical to pump design Multistage pump vibration can increase by a factor of 30 if the natural frequency of a shaft becomes equal to the frequency of a strong excitation force, such as residual imbalance at running speed, because this results in a resonance (Ref 2) Therefore, as a minimum, the designer must ensure that natural frequencies do not change sufficiently to cause a resonance within a range of clearance change of up to a factor of three Otherwise, shaft vibration will exceed the running clearances, and rapid wear will result

This can pose a significant problem for flexible shaft designs with a large number of stages, because the natural frequencies of such rotors depend strongly on the clearance of the annular seals through the workings of the "Lomakin effect" (Ref 3, 4) This effect causes a nonuniform pressure buildup through leakage velocity variation, as expressed by Bernoulli's equation This tends to recenter the shaft if it moves off-center and therefore contributes to the supporting stiffness and thus the natural frequency of the rotor The Lomakin in effect is illustrated in Fig 2

Fig 2 Schematic showing principle of the Lomakin effect (a) Longitudinal view of cross section (b) End view

Even if the designer has been careful to avoid vibration-related wear problems, pump life can still be dramatically reduced unless the user (1) runs the pump within the proper flow range and (2) monitors pump vibration The first concern is illustrated by Fig 3, which shows a typical plot of vibration versus flow rate Although common sense suggests that the less flow required of a pump, the less it will vibrate and wear, Fig 3 shows that the opposite is true The angles of the impeller and diffuser vanes of a pump can be designed to match the flow inlet and exit angles at only one particular flow rate, called the best efficiency point (BEP)

Fig 3 Dependence of pump vibration on flow rate

The user should select a pump whose BEP is close to the expected flow requirement of a particular system, because other flow rates, higher or lower, lead to a mismatch in flow-to-vane angles and result in potentially large steady and oscillating radial loads on the pump rotor (Ref 5) If the flow mismatch becomes great enough, stalling can occur (as on an airplane wing with too high an angle of attack), leading to formation of regions of recirculating flow and an associated increase of shaking forces (Fig 3) This is known as onset of recirculation and usually occurs first at the pump inlet or "suction" (Ref

6, 7) The minimum continuous flow of a pump should be set to avoid operating at or below this flow for extended periods of time This sometimes requires installation of a minimum flow recirculation loop, with a valve that automatically opens when flow drops below the recirculation onset flow

Another approach that is less wasteful than discharge throttling is to reduce pump flow as required by using a frequency drive (VFD) on the pump motor to reduce operating speed If VFDs are used, careful design of the support structures for vertically mounted pumps is essential to avoid resonances between the speed range and the structural

variable-natural frequencies, several of which are likely to be in the speed range (Ref 8) If throttling rather than VFDs is used, other vibration problems associated with the recirculation phenomena discussed above are possible (Ref 9)

It is also important that the user monitor pump vibration levels, allowing early identification and repair of vibration problems that may accelerate wear Various "condition monitoring" and "predictive maintenance" systems are marketed for this purpose Figure 4 gives limits of overall vibration as a function of running speed (Ref 10) Vibration above these limits can be expected to accelerate wear at a rate such that vibration at double these limits will result in up to 10 times the rate of wear Halving these limits generally has little or no effect, however, because at these levels no running clearance contact should have yet occurred; therefore, reducing vibration further has no impact on rubbing wear (Ref 1) Experience suggests that contact can be expected to begin, on average, at about one and a half to two times the recommended vibration limits

Fig 4 Allowable vibration as a function of running speed Adapted from Ref 10

Despite the best efforts of the designer and the user, it is likely that the occasion will arise when close clearances contact and rub Of key importance when such contact takes place is that no galling wear occur, because this can seize the pump and cause severe damage to the rotor system or drive train by overtorque

Galling can be resisted in several ways Opposing material pairs can be selected from "adhesive compatibility charts," such as the one shown in Table 1, which identify those combinations with the least chemical attraction This selection process is aided by heat treating the opposing parts so that they have a Brinell hardness difference of at least 50, which encourages wear of the softer material rather than adhesion and resultant part-to-part material transfer (Ref 11)

Table 1 Galling resistance of various material combinations

A second method of discouraging galling is to machine grooves in one or both of the close-clearance components, so that

as wear takes place the debris can collect somewhere other than at the close running clearance This also promotes rapid heat transfer at rubbing interfaces, keeping parts cool and hard Local surface temperatures can become very high even with grooving, because of the flash-temperature effect, but such temperatures decay in short distances and do not result in

galling if surface heat removal is effective (Ref 12) Grooving the surfaces results in a design compromise, however Although grooving reduces clearance leakage, it also reduces the beneficial shaft support provided by the Lomakin effect

Another way to assess galling potential is with a measure number for the contact stress required for cold welding and subsequent material pullout for a material mated against itself This is called galling threshold load and is tabulated in reference handbooks for various engineering materials Generally, austenitic and precipitation-hardening steels have very low galling threshold loads, on the order of 15 MPa (2 ksi) On the other end of the scale is Nitronic 50, Nitronic 60, and Waukesha 88, with values of about 350 MPa (50 ksi)

In open-impeller (that is, no shroud encloses the blade passages) and axial flow pumps, such as those used in water-flood service, the impeller or propeller blades interface directly with the casing across a narrow blade-tip clearance instead of with wearing rings The same considerations apply regarding use of compatible materials with reasonable hardness differences For large-diameter propellers, materials should also have relatively low friction coefficients, so that sudden rubbing contact does not cause overtorque of the pump shaft or drive train Materials selected using the compatibility chart (Table 1) generally have low friction coefficients Typical performance loss due to blade tip wear in unshrouded impellers and propellers is about 1% efficiency and head loss for each percent of blade height removal Open or semiopen impellers are often used for abrasive slurry service, particularly in the paper industry, as discussed below in the section on particulate erosion

Coupling tooth or shaft spline wear is typically a reliability problem only if the teeth are not properly lubricated or if they are overloaded by a vibration resonance (Ref 13) If resonance of the shaft occurs for a rotor natural frequency that includes significant coupling motion, reciprocating slip may occur at the coupling teeth, leading to fretting wear The only solution in this circumstance is to redesign the shaft or to use a different weight coupling to shift the natural frequency out

of running speed range

Bearing wear in pumps can take the form of gross plastic deformation, frictional softening or melting, and subsequent abrasive wear if there is a severe shaft overload or inadequate or improper lubrication Improper lubrication in such instances is usually caused by lubricant emulsification when contaminated with water by a seal failure or by atmospheric condensation More common than catastrophic bearing failure, however, is the gradual deterioration of the bearing by surface fatigue due to:

• Marginal lubrication, again perhaps because of lubricant contamination

• Steady loading somewhat in excess of the rated surface pressure of a journal bearing (between 1.4 and

14 MPa, or 0.2 and 2 ksi, depending on the bearing material and geometrical details) or above the rated 20,000 h B10 life of a rolling-element bearing

• Large oscillating loads caused by excessive vibration, usually because of either a rotor dynamic or a bearing housing natural frequency resonance (Ref 14)

In its final stages, surface fatigue is recognizable by loaded surface "craze cracking" and spalling of surface material in laminar chunks

Vertical turbine pumps (VTPs), used for deep well, low suction head, and wet pit service, are unique with respect to their lightly loaded bearings (gravity is parallel to the shaft) and the multiple-bearing support of the often long "line-shafting" that connects the deeply submerged impeller or bowl assembly to the above-ground motor driving the pump Another unique feature of VTPs is the "product lubrication" (that is, lubrication by the pumped fluid) of the lineshaft and bowl bearings Bronze or bronze-filled Teflon is a good bore material in such bearings, and their wear rate has been found to be

minimized by using a relatively large length-to-diameter ratio (an L/D ratio of 2.5 is common) The greater length of these

bearings tends to restrict the flexure of the relatively thin lineshafting, keeping whirling amplitudes and thus bearing loads low (Ref 15) In poorly lubricating fluids, such as water, and in abrasive fluids, rubber has been found superior to bronze, particularly when combined with axial or spiral grooves to feed bearing surfaces and wash out entrapped particles The inside surfaces of such bearings can be made polygonal or corrugated to further enhance lubricant feed and flush in the presence of low bearing loads This style of bearing, inexpensiveand popular in irrigation pumps, gives good service as long as the pump is not allowed to run dry

Reciprocating pumps are more prone to rubbing wear than centrifugal pumps, because they have several sliding components, including the following (Fig 5):

• Lower crosshead versus the crosshead ways

• Tie rods versus their linear bearing surfaces

• Plunger or piston rod versus the packing

• Plunger or piston surface versus the bore walls

• Valve lateral surfaces within the valve retainer or guide

In addition, each suction and discharge valve is subject to impact wear at the seating surface, particularly if cavitation is occurring Finally, there are several sets of highly loaded bearings: crankshaft bearings (one for each one or two cylinders plus the coupling), connecting rod "small-end" bearings attached to each lower crosshead pin, and connecting rod "large-end" bearings that attach each rod to the crankshaft

Fig 5 Cutaway view of a reciprocating pump Courtesy of Dresser Pump Division, Dresser Industries

To minimize wear in reciprocating sliding components, it is critical to provide clean flushing fluid to unlubricated areas and to filter the oil and rigorously maintain its quality for lubricated areas Ceramic plungers and bore liners together with clean flushing fluid are effective for avoiding excessive wear induced by third bodies during abrasive service In the crosshead ways, excellent alignment must be maintained with the crosshead to avoid locally high bearing pressures, and area-averaged surface bearing pressure during the highest load portion of the stroke should be kept well under 690 kPa (100 psi) by properly sizing the crosshead bearing surface during design

Valve surfaces and their retainers must be maintained at a high hardness (on the order of 40 HRC), with a hardness differential of 50 HB, and must be properly selected from the compatibility chart (Table 1) In addition, valves should be designed and tested such that absolutely no cavitation occurs during the initial valve opening, when static pressure is at its lowest Otherwise, severe valve impact or "hammer" can occur, damaging the valve and the seat and, through the transmitted elastic wave, jeopardizing the entire drive train (Ref 16)

Finally, special vertically mounted spherical or conical seat valves should be used in slurry or abrasive pumping service

or if deposits are expected to build up near the cylinders Such valves tend to discourage the trapping of debris in the seating interface (Fig 6) (Ref 17)

Fig 6 Cross-sectional view of abrasive-resistant slurry valves used in reciprocating pumps (a) Rib-supported

conical type (b) Spherical type Source: Ref 17

Reciprocating pump journal bearings tend to be highly loaded, with surface pressures from 6.9 to 13.8 MPa (1 to 2 ksi) Special bearing babbitt materials are needed, and configurations adapted to reciprocating motion, such as "bedded arc" nonsymmetrical bore designs for connecting rod small-end bearings, are used for higher load capacity (Ref 18)

It is also important that the bearing bore and journal be kept aligned and that the journal bore clearances be maintained This requires careful assembly, as well as a design procedure that accounts for transient thermal differences in the bore versus the journal and for elastic deformation of the bearing supports For example, if connecting rod "eye" cross sections

at either the small end or large end are not sufficient, the bearing can ovalize to the extent that all of its side clearance is taken up, "pinching" the journal and causing possible seizure of the bearing (Fig 7)

Fig 7 Bearing "pinch" generated by deformation of a small-end connecting rod

The crankshaft bearing is also prone to elastic support problems If the crankshaft is relatively stiff, care must be taken that the bearing support stiffness of the pump frame is optimized to carry the load and yet allow the bearings to translate and thus distribute the plunger loads of the rigid crankshaft to several bearings If the crankshaft is relatively flexible, on the other hand, the bearing support stiffness must be adjusted to avoid locally high bearing pressure at bearing corners when the crankshaft tilts within the bearing shell

Particulate Erosion

Problems of third-body erosion can occur in all types of pumps if they are subjected to abrasive service, such as slurry service Commonly pumped slurries include coal, mined phosphate, mined metal ore, and water with fly ash captured during flue-gas desulfurization The pumping of paper pulp is another common erosive service Such services must compromise hydraulic performance for wear resistance and require designs that allow ease of inspection and repair (Ref 19)

The abrasivity of a fluid can be quantified using the Miller number index, as outlined in ASTM Standard G 75 (Ref 20) The effective viscosity of an abrasive slurry is often quite high compared with the virgin fluid and should be considered along with abrasivity in selecting an appropriate pump Otherwise, the pump may have a head and flow capability that are too low for the intended service

Centrifugal Pumps. The typical centrifugal slurry pump is of an overhung single-stage end-suction design It should have reasonably low required value of net positive suction head (NPSH), as explained below in the section on cavitation This is because suction piping velocities must often be kept high so that the slurry stays above its "critical carrying velocity" or "nonsettling velocity" below which the slurry cannot be kept evenly suspended (Ref 21) This requirement is particularly strong if the slurry particles are larger than roughly 65 m

The areas of a centrifugal pump most prone to abrasive wear are the impellers and the casing interior, where flow-path velocities are high One approach to decreasing erosive wear of these areas and to maintaining the pump is the use of hard, wear-resistant, replaceable liners Elastomeric liners are sometimes used as an alternative to hard metal liners Natural rubber, neoprene, nitrile, and chlorosulfonated polyethylene have exhibited high abrasion resistance when used as external liners for impellers and as internal liners for casing volute passages However, care must be taken that the pressure in the flow passage next to the liner does not fall below the pressure between the liner and the wall, which is typically close to atmospheric pressure When this occurs in service, unless the liner is mechanically bolted to the casing

at key locations, it may billow out, catching the impeller blades and destroying the rotor Another caveat is to stay below the softening temperature of the rubber, typically about 115 °C (240 °F) Also, for elastomeric materials to exhibit good abrasion resistance, the abrasive particles must be less than 5 mm (0.2 in.) if they are dull and less than 2 mm (0.08 in.) if they are sharp Finally, the head per stage should be kept below about 50 m (165 ft) (Ref 19)

When metal impellers and casing liners are used, they should be as hard as possible, not only because wear resistance generally increases with hardness, but also because abrasive erosion in pump components generally occurs due to the impingement of particles at very oblique angles to the surface As pointed out by Finnie (Ref 22), such abrasion attacks ductile materials much more readily than it does hard, relatively brittle materials Finnie's equation for hard materials is:

Wear rate = (Number of impinging particles) · (Average particle mass) · (Impingement velocity)2 · (Angle of impingement)

(Eq 1)

Particulate erosion of a ductile steel impeller by a catalytic fluid that was not expected to be erosive is shown in Fig 8

Fig 8 Particulate erosion of a ductile steel impeller by an abrasive catalyst (a) 0.25× (b) 1×

Typical hard, erosion-resistant metals include white cast iron (standard gray or ductile cast irons are poor), chromium (that is, 13 to 28%) alloy steel, cobalt-base superalloys such as Stellite, and special nickel-base alloys such as Ni-Hard These materials are useful not only for the flow-path surfaces but also as sleeves in sealing areas, particularly if packing is used It is good practice to use hard metal or ceramic expellers near seal entrances to keep particles out of the sealing clearance If thorough flushing of packing presents a design problem, abrasion-resistant SiC ceramic mechanical seals together with an expeller or minimal flushing can be used Some of the newest designs of mechanical seals for abrasive service are even axially split, for ease of replacement

high-Design steps other than material selection can be taken to maximize the ability of a pump to withstand abrasives For example, the casing volute tongue can be retracted to avoid the highest-velocity region of discharge flow, and casing volute walls can be contoured to minimize impingement angles in high-density zones of particle trajectories (Ref 23) The impeller can be redesigned with its blades recessed into toroidal cavities, thus avoiding many mainstream particles (however, a stiff efficiency penalty can be expected for this configuration) Impeller vanes can be made thicker so that it takes longer for their edges to erode Also, larger pumps than required can be selected and then run at less than design speed, through use of a simple adjustable pulley or variable-frequency drive excitation of the motor (Ref 24)

Even if optimum materials and design configurations are chosen, centrifugal pumps still face practical limitations in abrasive service For example, head per stage should be kept below 100 m (330 ft) or else the implied volute velocities will be too high When higher heads are required, pumps can be arranged in series Also, abrasive particles may collect in impeller side clearances, held in suspension by the counter-tendencies of the leakage flow, which is biased to moving toward lower diameter by design, and of centrifugal forces, which always work away from minimum diameter This can result in long-lived, particle-filled vortices (Ref 25), which quickly erode the impeller hub unless hardened wear plates are applied at the proper locations

Open or semiopen impellers are often used for abrasive or slurry service, particularly in the paper industry Such impellers can be designed with clearance in the axial direction, which sacrifices new pump efficiency but makes the pump performance less prone to clearance increases and allows clearance adjustment by shimming of the impeller to change its position on the shaft The surfaces of the impellers and casing flow paths, particularly the volute, must be made of a hard metal or ceramic or of a tough elastomer, as discussed above

Reciprocating and Other Positive-Displacement Pumps. Generally, centrifugal pumps are the best choice if relatively low heads and high flows are required; otherwise, reciprocating pumps are the best choice Other criteria, however, are involved in abrasive service As discussed above, centrifugal pump heads face certain limitations in such applications On the other hand, for particle sizes greater than about 2.5 mm (0.1 in.), even the most particle-tolerant reciprocating pump valves will tend to jam and are not satisfactory

In such cases, a diaphragm pump driven by air or clean fluid may be suitable up to about 100 m (330 ft) of developed head Above this pressure, a progressive cavity pump might be considered (Ref 26) One type of progressive cavity pump, the single- or double-screw pump shown in Fig 9, is capable of high pressure and relatively high flow In abrasive service, screw pumps have been shown to pump abrasives and even mixed-phase fluids reliably with the use of hard metal screws and high-durometer (about Shore 68) rubber liners

Fig 9 Cross-sectional view of a double-screw pump (a) Top view (b) Side view Courtesy of Dresser Pump

Division, Dresser Industries

If particle size is below 2.5 mm (0.1 in.), reciprocating pumps are applicable, and in services where they compete with centrifugal pumps, they may even be the economical choice because of their high thermodynamic efficiency (on the order

of 95%) (Ref 27) Two classes of reciprocating pumps are piston and plunger types The piston types are generally superior for higher flows, but the plunger types can better tolerate sharp and highly abrasive particles

Special vertically mounted spherical seat or reinforced conical seat valves should be used in slurry or abrasive pumping service, or if deposits are expected to build up near the cylinders Such valves tend to discourage the trapping of debris in the seating surface, as illustrated by the valves shown in the earlier discussion of rubbing wear (Fig 6) (Ref 17)

Cavitation Erosion

Cavitation occurs in portions of the flow path where static pressure drops below the vapor pressure of the pumped fluid This causes very small vapor bubbles to form, which then rapidly and violently collapse when they enter an area where static pressure once again exceeds vapor pressure The sudden bubble collapse causes concentrated shock waves that impinge on nearby metal, resulting in stresses large enough to induce fatigue

Unlike surface fatigue caused by rubbing, which produces subsurface cracks that eventually delaminate the surface in thin flakes, many cavitation fatigue cracks form perpendicular to the surface and tend to remove material in small blocks The resulting damage leaves a surface finish that initially is matte, later becomes pock marked, and eventually has the appearance of rough, mountainous terrain, with odd eroded surface protrusions Cavitation is probably the most common cause of rapid component erosion and wear in hydraulic machinery

The usual measure of likelihood of cavitation is the net positive suction head (NPSH), which is the difference between the static pressure and the vapor pressure of the liquid at the pump inlet flange, expressed in meters:

(Eq 2)

where Ps is the static pressure at the pump inlet flange (in Pa), Pv is the vapor pressure of the liquid (in Pa), is the liquid density (in kg/m3), and g is the gravitational constant, 9.8 m/s2 A required value of this number, the NPSHR, is determined by the pump manufacturer, generally on the basis of when sufficient bubble formation occurs to drop the head produced by the pump by 3% The system's available NPSH, or NPSHA, must be at least equal to this value If it is not, either another pump that is less demanding at the suction must be chosen or some means must be found to increase the NPSHA, such as an upstream booster pump or, in a centrifugal pump, use of an axial inducer The latter looks like a corkscrew protruding from the first-stage inlet and controls the inlet flow angle while adding to the initial suction pressure Care must be taken to include in NPSHA the losses due to suction pipe friction, valves, and strainers In reciprocating pumps, the effect of pulsating flow inertia, or "acceleration," can reduce transient suction pressure dramatically and must be accounted for and reduced, if necessary, by using a pulsation dampener

The situation became complicated in recent years when researchers found that NPSHR based on 3% head drop was insufficient to avoid cavitation damage Instances occurred in which more than three times as much suction head was required to avoid bubble formation and subsequent damage Other more stringent criteria have been proposed, such as

"1%" and "0%" (that is, on the verge of total head drop) (see Ref 28), but the pump industry still generally uses the familiar 3%, with a factor of safety based on experience for a given class of pumps Some researchers believe that observation of cavitation beyond a certain bubble length should be the criterion (Ref 29), but so far, general application of this has been impractical

An early attempt to predict cavitation used the Thoma number, which is a ratio of NPSHA to total head This number, originally applied to hydraulic turbines, places too much emphasis on discharge conditions to be generally applicable to

pumps A modification is the cavitation number or coefficient, K, which is a ratio of NPSHA to suction tip speed squared

K is referenced to the suction, but does not account for local increases in velocity and associated decreases in static

pressure in the flow passages, especially in the turbulent and sometimes partially stalled inlet of impellers

Another factor of importance is the ratio of NPSHR NPSHA to vapor pressure, where NPSHR is the value of NPSH at which cavitation damage begins (not necessarily the same as the 3% NPSHR, as discussed above) Once cavitation damage begins, this ratio is higher for cold fluids than for warm ones, because a warm fluid has a higher vapor pressure than a cold one This higher ratio suggests that cavitation is more severe in cold fluids, which is exactly the case A warm fluid, with higher vapor pressure, will cavitate at a lower suction pressure; however, the rate of bubble collapse and hence the severity of the cavitation once it occurs is roughly dependent on the ratio of the static pressure differential to the vapor pressure, so the bubble collapse damage is less severe

Centrifugal Pumps. A useful approach for assessing the potential for cavitation damage in centrifugal pumps is the

concept of suction-specific speed, S, developed by Karassik (Ref 30):

(Eq 3)

As S becomes larger, the chances become greater that cavitation damage will occur despite the 3% head drop NPSHR value being met, and if it does occur it is more severe, on average, as S increases Cavitation in pumps without inducers becomes highly probable as S increases beyond the range of 13,300 to 16,000, where NPSH is in feet and flow is in

gallons per minute (Ref 31)

An indirect but important cause of cavitation is suction and discharge recirculation (Ref 6, 7), which occurs when the pump is run well below the flow it was designed for, and the inlet or exit flow angles match poorly with the pump vane angles This leads to local secondary flow eddies on the high-pressure side of the blade The locally increased velocity in these eddies can decrease local suction pressure sufficiently to cause cavitation and subsequent damage This damage is usually distinguishable from direct cavitation, because it usually occurs on the pressure side of the blade Direct cavitation generally occurs where the flow-path static pressure is low in the primary flow namely, on the suction surface of the blades, usually close to the inlet An example of direct cavitation damage is shown in Fig 10

Fig 10 Early stages of centrifugal pump cavitation and damage

As discussed earlier in the section on rubbing wear, recirculation and its associated problems can be avoided by ensuring that the continuous (that is, long-term) flow of the system is not below the flow associated with the onset of suction recirculation, which is typically somewhere between 25 and 75% of the rated capacity of the pump This may be difficult

in an existing installation, and a redesigned impeller to lower recirculation onset flow should be considered, although this may reduce efficiency somewhat

In instances of either direct or recirculation-induced cavitation damage, if it is not possible to directly fix the cause (that

is, by increased suction pressure or increased minimum flow, respectively), then a material with greater cavitation erosion resistance can be used to replace the existing impellers In general, cast iron, low-carbon steel, and austenitic steels have low cavitation resistance, although they are often used where significant cavitation is not expected Titanium, aluminum bronze, Nitronic 50, and a new alloy called Hydrolloy have good cavitation erosion resistance, the tradeoff being increased expense

A standardized accelerated test procedure (ASTM G 32) (Ref 32) is available for assessing the cavitation damage resistance of materials if existing data cannot be found This procedure is based on the use of high-frequency vibrations to excite cavitation close to the metal surface

Reciprocating Pumps. The components of a reciprocating pump most likely to suffer cavitation damage are the suction valves Suction valves should be designed and tested for a given pump flow and quoted NPSHR such that absolutely no cavitation occurs during the initial valve opening, when the annular valve area is small and fluid velocity is high, causing locally low pressures through Bernoulli's equation When even a small amount of cavitation occurs in one valve, it triggers an "impulse response" of the suction manifold acoustic natural frequencies, with resulting pressure pulsations large enough to make all of the suction valves severely cavitate This can cause "valve hammer," where the valve sharply impacts its seat at an acceleration in excess of 100 G's The resulting shock tends to peen the valve seating area, causing fatigue, and can send strong stress waves through the entire drive train This can loosen bolted connections and cause fatigue or bearing problems far from the source of the problem (Ref 16)

If suction valve cavitation occurs, either suction pressure must be increased or the valve design must be changed to lower the NPSHR When determining reciprocating pump suction pressure, it is important to include the suction line inertia effects when the suction valve opens, known as acceleration head This reduces the effective NPSHA and can be minimized by shortening the suction piping or by placing a large-volume pulsation dampener in the piping, as close to the pump as possible (Ref 17)

Additional information about cavitation damage can be found in Ref 33 and in the article "Cavitation Erosion" in this Volume

References

1 W Marscher, "The Relationship Between Pump Rotor System Tribology and Appropriate Vibration Specifications for Centrifugal Pumps," Paper C123/87, Institution of Mechanical Engineers, 1987

2 W Marscher, Structural Design and Analysis, Sawyer's Gas Turbine Handbook, D Japikse, Ed.,

Turbomachinery International Publications, 1985

3 H Black and D.N Jenssen, Dynamic Hybrid Properties of Annular Pressure Seals, Proc J Mech Eng., Vol

7 B Schiavello and M Sen, On the Prediction of the Reverse Flow Onset at the Centrifugal Pump Inlet,

Proceedings of the ASME Symposium on Performance Prediction of Centrifugal Pumps and Compressors,

American Society of Mechanical Engineers, 1980

8 W Marscher, Reliability of Vertical Pumps, Proceedings of Water and Wastewater Conference '90

(Barcelona), 1990

9 W Marscher, "Subsynchronous Vibration in Boiler Feed Pumps Due to Stable Response to Hydraulic Forces at Part Load," Paper C349/88, Institution of Mechanical Engineers, 1988

10 API-610 Pump Specification, 7th ed., American Institute of Petroleum Engineers, 1989

11 W Marscher, a Phenomenological Model of Abradable Wear in High Performance Turbomachinery, Wear,

14 W Marscher, Vibration Test and Analysis of a Barrel Boiler Feed Pump Exhibiting Non-synchronous

Vibration, Proceedings of IMechE Seminar on Vibration in Centrifugal Pumps, Institution of Mechanical

Engineers, 1990

15 W Marscher, The Effect of Fluid Forces at Various Operation Conditions on the Vibrations of Vertical

Turbine Pumps, Proceedings of IMechE Seminar on Pumping, Institution of Mechanical Engineers, 1986

16 S Collier, Know Your Triplex Mud Pump, Parts 1-6, World Oil Mag., Jan-June 1982

17 J Miller, "Reciprocating Pumps for Slurry Service," ASLE Preprint 84-AM-6A-1, American Society of Lubrication Engineers, 1984

18 J Campbell et al., Bearings for Reciprocating Machinery: A Review of the Present State of Theoretical, Experimental, and Service Knowledge, Proceedings of the IMechE Conference on Lubrication and Wear,

Institution of Mechanical Engineers, 1968

19 J Doolin et al., Pumping Difficult Fluids, Chem Eng., Dec 1991, p 67-79

20 "Test Method for Slurry Abrasivity by Miller Number," G 75, Annual Book of ASTM Standards, Vol 3.02,

ASTM, 1984, p 420-437

21 G.R Addie, "Centrifugal Slurry Pump Selection and Application Tutorial," presented at 4th International Pump Symposium, Texas A&M University, 1987

22 I Finnie, Some Observations on the Erosion of Ductile Metals, Wear, Vol 19, 1972, p 81-90

23 M.C Roco and G.R Addie, Analytical Model and Experimental Studies on Slurry Flow and Erosion Flow

and Erosion in Pump Casings, Proceedings of the Slurry Transfer Association, March 1981

24 W O'Keefe, Pumps, Valves, and Piping, Power Mag., March 1992, p 19-30

25 A.P Smith, Unexplained Wear in Large Centrifugal Pumps, Chem Eng., Aug 1979, p 153-155

26 A Prang, "Rotary Screw Pumps for Multiphase Products," presented at Conference on Seals and Vibration Reliability of Centrifugal Turbomachinery (Ukraine), 17-20 Sept 1991

27 W Smith, Construction of Solids Handling Displacement Pumps, Power Fluids Mag., Vol 9 (No 1), 1983

28 E Grist, "Net Positive Suction Head Requirements for Avoidance of Unacceptable Cavitation Erosion in Centrifugal Pumps," Paper C163/74, Institution of Mechanical Engineers, 1974

29 J Gulich and A Rosch, Cavitation-Erosion in Centrifugal Pumps, Sulzer Tech Rev., No 1, 1988, p 28-32

30 I Karassik, Centrifugal Pumps and System Hydraulics, Chem Eng., Oct 1982, p 84-106

31 J Hallam, Centrifugal Pump Suction Specific Speed, How Important Is It?, Proceedings of the ASME Petroleum Mechanical Engineering Conference, American Society of Mechanical Engineers, 1982

32 "Method for Vibratory Cavitation Erosion Test," G 32, Annual Book of ASTM Standards, Vol 3.02, ASTM,

1988

33 J Doolin, Judge Relative Cavitation Peril with Aid of These Eight Factors, Power Mag., Oct 1985

Friction and Wear of Compressors

Royce N Brown, Dow Chemical Company

Introduction

A COMPRESSOR is a mechanical device that is used to increase the pressure level of compressible fluid These pressures can be at any level, from vacuum to high positive levels The fluids can consist of either vapors or gases, singularly or as mixtures Because the compressors described in this article use mechanical motion to perform the compression function, they are subject to friction and wear Each of these compressors is defined below by type and function, followed by a brief discussion of how friction and wear are manifest in each

Fig 1 Reciprocating compressor cycle

The distinguishing feature of the trunk-type compressor (Fig 2) is that the piston is directly connected to the connecting rod This type of construction is normally used in smaller machines, such as small shop air compressors and refrigeration service compressors

Fig 2 Trunk-type reciprocating compressor Courtesy of Ingersol Rand

The crosshead-type compressor (Fig 3) is normally found in process gas service The connecting rod is attached to the crosshead, which is connected to the piston rod that moves the piston The cylinders are normally double acting, a feature that permits compression to occur on either side of the piston, which essentially doubles the capacity for a given-sized cylinder It can be noted that a gas seal in the form of packing can prevent gas leakage from the compressor cylinder to the crankcase

Fig 3 Cross section of crosshead-type reciprocating compressor Source: Dresser-Rand

The cylinder, which is the gas-handling part of the reciprocating compressor, contains the piston, which is the gas displacer Motion is transmitted to the piston by the connecting rod in the trunk-type compressor, and by the piston rod in the crosshead type The cylinder includes the valves that control the inlet and outlet gas flow The compressor valves are spring loaded and gas actuated There are one or more inlet valves and one or more discharge valves Flow and pressure considerations would dictate the number and when multiple valves are needed

Cylinder Wear. The piston includes a set of seals, more commonly referred to as piston rings Because the rings rub the cylinder walls, wear of both the rings and the cylinder wall becomes a consideration in material selection The majority of cylinders used for process gas service utilize replaceable liners that act as the cylinder wear surface

Cylinder orientation has an influence on the wear of the cylinder and rings Vertically oriented cylinders would be preferred if wear were the only consideration Small compressors are generally constructed with the cylinders vertical, or nearly so in a "V" configuration As compressor size increases, height also increases, which causes head-room problems and makes maintenance unwieldy

Generally, the larger compressors are arranged with horizontal cylinders; this arrangement makes spaces management easier and facilitates maintenance However, these gains are somewhat offset by the effects of gravity The weight of the piston causes additional rubbing forces to the cylinder wall and increases the potential for wear, particularly to the rings

To counter this, nonmetallic (usually) rider rings are supplied with the pistons The rider rings are larger and act as the

wear surface for the piston American Petroleum Institute (API) Standard 618 provides guidelines on the loading limits of rider rings used in reciprocating compressors for refinery service

As stated earlier, the piston rings act as a seal to minimize the quantity of gas that can bypass the piston Again, the need for compromise becomes evident If there are many rings and/or a very tight fit to the cylinder wall, then sealing is excellent, but friction force and wear are increased Besides the wear considerations, increased friction results in a larger drag on the piston This absorbs energy and results in an increased parasitic load on the compressor Ultimately, this loss results in a decrease in efficiency A piston is shown in Fig 4

Fig 4 Crosshead-type reciprocating compressor piston and rod Courtesy of Dresser-Rand

Piston speed is a key factor in the wear of a cylinder and the rings Average piston speed is defined as:

where PS is piston speed in m/s, S is stroke in m, and N is crankshaft speed in rev/s Piston speeds normally range from 2

to 6 m/s (400 to 1200 ft/min) It is obvious that the higher the piston speed, the higher the wear Industrial compressors are normally selected in the range from 3 to 4 m/s (600 to 800 ft/min) If the cylinder is of a nonlubricated design, then the piston speed should be toward the lower end of the given speed range to minimize wear

Cylinder lubrication depends on the type of compressor being considered In trunk-type compressors, cylinder lubrication is normally obtained from the crankcase, and reaches the cylinder wall by a splash action Although not sophisticated, this method does provide adequate lubrication

In crosshead-type compressors, cylinder lubrication is divided into either nonlubricated or lubricated cylinder compressors The nonlubricated cylinder compressor does not supply an outside source of lubricant to the cylinder The term nonlubricated is probably not the most accurate description of the lubrication process In actuality, the ring materials are chosen for their inherent low friction or self-lubricating characteristics The lubricated cylinder compressor is furnished, by an outside source, with a lubricant that is either a conventional mineral oil or a synthetic oil, depending on the process gas

Two methods of supplying the lubricant to the cylinder and rod packing are available The amount of lubricant furnished

to any lubrication point is quite small, on the order of droplets per minute Use of the metering-type lubricator permits precise tailoring of lubricant quantity to the cylinder points and packing without compromising the frame lubrication system, which is much larger

One type of mechanical lubricator consists of a multiplunger pump, where an individual plunger is dedicated to each lubrication point A sight glass is normally included in each plunger feed to allow visual monitoring of the point Another common type of mechanical lubricator system consists of a single metering pump sized for the total combined flow to all points A divider block distributes the flow in the required amount to the individual lubrication points Each system has

pros and cons associated with it Generally, the compressor supplier can provide recommendations to users who have not developed a preference

Cylinder Materials. In many smaller compressors, the wearing portion of the cylinder is made of the same material as the balance of the cylinder (usually, gray cast iron) In some instances, a cylinder can be nodular, or ductile, iron On smaller high-pressure service cylinders, steel is used Smaller compressors of the more standardized type normally use an unlined cylinder

On larger compressors, cylinders with replaceable liners are more common Liner material can be chosen for its wear properties, as well as compatibility with the application If cylinder wear does occur, then the liner construction represents

a renewable part For an unlined cylinder, the only recourse after significant cylinder wear is to either rebore or replace the cylinder

The piston ring materials are made of either a fluorocarbon compound or polyether etherketone (PEEK) for nonlubricated service The fluorocarbon is generally a carbon-filled type For lubricated service, rings can be either metallic or nonmetallic Bronze and cast iron are common for metallic service, whereas carbon-filled fluorocarbon is normally used for nonmetallic service Rider rings generally use carbon-filled fluorocarbon

Valves are used to control the gas that either enters or leaves a cylinder Valves can be one of the higher-maintenance items on the reciprocating compressor Because valve duty is quite severe, it is subject to cyclic stresses that are due to the requirement that it open and close on each stroke Upon each open and close cycle, the valve is subjected to impact loads

It also experiences all of the temperature changes that occur in the compressed gas On lubricated cylinders, the presence

of lubricant does not extend valve life Valve motion is difficult to lubricate Also, in some cases, the lubricant in combination with high temperatures may cause carbon to form on the valves, shortening their lives As the valves wear, compressor performance deteriorates, which lowers efficiency and raises temperature, furthering the deterioration problem

A plate valve is shown in Fig 5 and a channel valve is shown in Fig 6 In most cases, the material of construction is metallic Gray iron, nodular iron, and steel are commonly used for valve seats and guards Valve springs are chrome, silicon, stainless steel, or inconel Higher alloys may be needed for reactive process gases Valve plates are usually stainless steel, plastics, or PEEK

Fig 5 Plate valve Source: Dresser-Rand

Fig 6 Channel valve Courtesy of Dresser-Rand

Piston rod packing is required on double-acting cylinders where the piston rod passes through the crank-end cylinder closure This may also be true on the head end if a tail rod is used The purpose of the packing is to contain the gas in the cylinder The piston rod must have a high degree of finish in the packing area Various methods of surface hardening are used These features are needed to minimize the packing and rod wear in the packing region On lubricated compressors, the packing is given lubrication from the cylinder lubricator, which helps reduce the friction and wear at this point The packing rings are generally a segmental type, and are usually made from a carbon-filled fluorocarbon-type material

In addition to being a source of wear, packing is another source of friction, which can contribute to loss in efficiency Gas release that is due to excessive leakage is an extreme result of excess wear

The crosshead is the point where the connecting rod motion is converted to pure linear motion It provides the feature that makes double-acting cylinders possible A crosshead and associated parts are shown in Fig 7 The crosshead is lubricated by the frame lubrication system The critical requirement for the crosshead is that the piston rod provide a force reversal at the crosshead pin This reversal causes the pin to change its position in the crosshead bearing clearance This movement allows oil to enter both sides of the bushing to ensure adequate lubrication

Fig 7 (a) Crossheads (b) Crosshead with pin and shoes Courtesy of Dresser-Rand

On a fully loaded cylinder, rod reversals are not a problem However, if one end of the double-acting cylinder is unloaded, the gas load could exceed the inertial load and prevent the reversals from taking place Clearance pocket unloaders or external clearance bottlers can be used for alternate unloading schemes to solve rod reversal problems

The frame portion of the reciprocating compressor is a major component of the compressor, in terms of physical size and weight It houses the crankshaft, where rotary motion is translated into the reciprocating motion required to move the pistons It provides the mounting base for the cylinders and provides for their alignment to the crankshaft, which is cradled in a set of main bearings located in the frame The crankshaft also has journals for the connecting rod bearings

These bearings can be either the hydrodynamic oil film type or the rolling-element type There may also be a combination; the main bearings could be one type and the connecting rod journals, the other type Normally, the larger process compressors use the hydrodynamic type for all bearings

On the crosshead-type compressor, the crosshead housing is attached to the basic frame member, and the distance piece and cylinder, in turn, attach to the crosshead housing In some designs, the crosshead housing is integral to the basic frame member

It could be said that the frame makes a contribution to valve life This connection is not really all that mysterious, in that the crankshaft speed dictates the valve cycle rate If the crankshaft speed is high for a given service, then the valve cycle rate is high The cycle rate has a direct impact on valve life

Frame Lubrication. The frame section of the reciprocating compressor is independently lubricated in crosshead-type machines For the trunk type, it has been noted that all lubrication, both for frame and cylinders, is supplied by the frame system One method used in smaller machines is a very simple nonpressurized system, which lubricates by splashing the oil The splashing is accomplished by the dipping of the crank throws into oil contained in a sump in the crankcase Dippers can be attached to the connecting rods to aid in the oil distribution

Pressurized systems have at least one oil pump, which is driven by either the compressor's crankshaft or a separate electric motor There can be additional pumps, such as a start-up pump for shaft-driven main pump configurations The auxiliary pump may alternately take the form of a full-sized electric motor standby A filter that provides for oil cleanup can be duplicated to allow servicing while the compressor is operating

When the heat duty cannot be dissipated in the sump, an external oil cooler is required This exchanger is either a cooled or air-cooled device For critical service, two coolers can be used to permit servicing while the unit is operating Service valves and instrumentation for monitoring pressure, temperature, and level also may be included depending on the critical nature of the service and on the size and type of unit

liquid-Reciprocating Compressor Overview. The interaction of economics and life is an ever-present challenge to the user

of any equipment, but this is particularly true for the reciprocating compressor If, for example, a volumetric rate of gas is specified, then several choices are available First, a cylinder size and speed are selected, and then a small-diameter long stroke can be chosen This has some first-cost advantage over choosing a larger bore, shorter stroke The disadvantage is that the piston speed may be too high for good life Crankshaft speed is also a factor A higher speed will reduce the overall size of the cylinder and again lower the first cost, but this impacts both piston speed and valve cycle rate

The final selection is normally the best compromise of first-cost and compressor component life This will vary from application to application, based on the critical nature of the service and the ability to withstand downtime Even if downtime is not considered serious, another factor is the projected cost of maintenance and its impact on the life cost of the machine

Rotary Compressors

Many rotary compressor types have a low-pressure design and are generally referred to as blowers Their pressure differential capacity is normally limited This discussion focuses on rotary compressors with a somewhat higher pressure range, as well as the capability for a higher pressure rise To some degree, these compressors are capable to competing with the reciprocating compressor

Like the reciprocating compressor, the rotary compressors are of the positive-displacement type In actuality, they supplement the range of applications of the reciprocating compressor, rather than duplicate it

Lobe Type

Lobe-type compressors all have some features in common However, the most significant type is the helical lobe, or screw compressor, which is also referred to as the SRM and Lysholm compressor It can be configured as either a non-flooded (dry) or flooded compressor The dry configuration is generally larger in capacity, but is limited to a lower pressure ratio The flooded compressor is used for lower volumetric flow, but can achieve a higher pressure ratio

The dry helical lobe compressor consists of two rotors, one male and one female, which are arranged in their casing to intermesh when they turn The compression cycle is shown in Fig 8 The gas path takes the form of the helix cut in the female rotor The rotors are timed by a set of timing gears located at one end of the machine Gas path sealing is accomplished by seal strips located at the tips of the rotors Two rotor profiles are shown in Fig 9

Fig 8 Compression cycle of helical lobe compressor

Fig 9 Rotor profiles

The rotor discharge ports are cut to match the application pressure ratio The purpose is to release the pocket of trapped gas being compressed at the application level to prevent overcompression and undercompression This allows the compressor to operate at its optimum efficiency A set of helical screw compressor rotors is shown in Fig 10

Fig 10 Helical screw compressor rotors

Rotor Wear. The rotors, in ideal service, would not wear because they would not touch, being synchronized by the timing gears Realistically, wear does occur Although this style of compressor works better in dirty gas service than many other compressors, the presence of particulates are what causes wear Normally, this occurs on the seal strips This compressor also tolerates liquid mist better than the other compressors If this type of operation is anticipated, the supplier can overlay the casing with a welded material to add abrasion or corrosion resistance and to minimize rotor bore wear Although seal strips can be renewed, it is difficult and therefore expensive to repair a badly eroded casing

Bearings and Seals. Dry helical lobe compressors normally have hydrodynamic radial and thrust bearings, which are pressure fed from an outside oil system Because of the nature of the oil film that develops from rotation, hydrodynamic bearings do not suffer from much wear Wear normally is due to the lack of lubricant cleanliness If the particle size in the lubricant is larger than the minimum lubricant film thickness, then excess wear occurs This is true for the four journal bearings and the thrust bearing A well-maintained lubrication system with properly sized oil filters should prevent this

There are a few applications for which rolling-element bearings may be used by the manufacturer However, this is not very common on dry helical lobe compressors used in process service

The dry helical lobe compressor requires four shaft end seals, which can have many configurations The application dictates the basic requirements of leakage On an air compressor, a relatively simple labyrinth or carbon ring seal is adequate This type of seal is generally referred to as a restrictive seal, because it limits, rather than stops, leakage Of the two, the carbon ring seal is more effective, but unfortunately is more subject to wear than the plain labyrinth

Two other basic forms of seals are commonly used, depending on application requirements: the liquid film and the mechanical contact seal The liquid film seal uses a metallic sealing ring, which is liquid buffered to maintain a liquid film in the ring clearance area The presence of the liquid film excludes the gas, thereby effecting a seal The liquid normally is a lubricant, which helps prevent wear to the bushing Water can be used for applications where the lubricant and gas are incompatible Carbon bushings are sometimes used in water-buffered service

The mechanical contact seal is more expensive, but also provides the greater degree of gas containment This seal type can have either a liquid buffered or dry gas form The preferred fluid for buffering is a lubricant When lubricating oils is compatible with the process, lubricant from the bearing lubrication system can be used The control of the lubrication system pressure becomes more complicated, because several pressure levels may be required However, no additional liquid supply system is needed In some rare applications, this seal is used with a liquid that has poor lubrication properties The result is a higher maintenance cost, because of wear

The dry gas seal is relatively new and applications in dry helical lobe compressors have been limited This type of mechanical contact seal uses a clean dry gas as a buffer The gas forms a separating film so that the seal does not rub The seal uses less-parasitic power than its liquid buffered counterpart It can be obtained in a variety of configurations To prevent premature failure that is due to excessive wear, the seal buffer gas must be well filtered

When the application requires very low leakage, the mechanical contact seal should be considered However, as of early

1992, no commercial seal with zero leakage is available for these compressors With careful engineering, leakage can be either contained and retrieved or disposed, in most cases

The flooded helical lobe compressor uses the same lobe form as its dry counterpart The significant physical difference is that it uses the flooding fluid to time the rotors, rather than a timing gear Rotor-to-rotor sealing is also accomplished by the flooding fluid

This type of lobe compressor is normally smaller, and is used mostly in lower-flow applications A benefit of flooding is the removal of the heat of compression by the flooding fluid A disadvantage is that the same flooding fluid is used to lubricate the bearings If the gas is not compatible with lubricating oil, then special bearings may be required to operate with a gas-compatible flooding fluid

Because of the size of the flooded compressor, rolling-element bearings are commonly used, although hydrodynamic bearings are also quite common Seals are of the buffered mechanical contact type, because the sealing fluid must be contained with the gas

An advantage of the flooded helical lobe compressor is that it is a relatively simple machine It can accommodate a moderately high pressure ratio in a single casing Therefore, it is more compact than some of its competitors Its major

disadvantage is due to the same feature that provides its advantage: the flooding fluid Incompatibility with the fluid limits the possible applications The fluid must be removed from the gas stream to prevent interference with the process This requires complex separation equipment in some applications Additionally, the auxiliary system to supply the lubrication and flooding fluid is quite large In many cases, it dwarfs the compressor

Sliding Vane

The sliding vane compressor is a positive-displacement rotary-type compressor It consists of a slotted rotor mounted eccentrically in a cylinder that is slightly larger than the rotor The rotor slots hold a set of vanes, which are free to move radially within the slots Centrifugal force moves the vanes out of their slots and keeps them in contact with the cylinder wall A cross section of this compressor is shown in Fig 11

Fig 11 Cross section of sliding vane compressor Courtesy of AC Compressor Corporation

The space between the rotor, set of vanes, and the cylinder forms the gas cell The inlet is located such that the cell expands when passing the inlet port The cell then decreases in size as the rotor continues to turn, compressing the gas The outlet port is strategically located to be passed when the compression matches the outlet conditions This compression process is similar to the one described for the lobe compressor

Components. The cylinder is most commonly made of cast iron, with the bore being machined to a fine finish The vanes are made of a laminated construction impregnated with a phenolic resin The vanes are in continuous rubbing contact with the cylinder wall, which causes vane wear This, of course, makes vane replacement a routine maintenance item

To minimize the effects of vane friction and wear, most of the rotary vane compressors inject a lubricant into the cylinder

A mechanical plunger-type metering pump is commonly used, and is similar to the point lubricator described for the lubricated reciprocating compressor Although there is much less lubricant than would be used in a flooded helical lobe compressor, it represents a disadvantage if there are gas or process compatibility problems Its removal further compounds the problem

Rolling-element bearings are commonly used in all but the largest of these compressors With the plunger-per-point lubricator, a point is normally devoted to each bearing, which takes care of bearing lubrication

Centrifugal Compressors

The centrifugal compressor normally is applied in applications where the volumetric flow is larger than that which is typical of positive-displacement compressors Compared to the latter, the pressure ratio across a typical centrifugal compressor is less

This compressor is different from the positive displacement type because it causes gas to be compressed by using gas dynamics The active element of the centrifugal compressor is the impeller All work done to the gas occurs in the impeller The balance of the gas path components are used to convert the energy in the gas from velocity head to pressure Other components are used to guide the gas through the compressor with minimum losses

The impeller consists of a set of vanes enclosed in a front and back shroud, as in an enclosed design The front shroud can

be stationary, with the vanes attached to the back shroud, as in a semienclosed design Sometimes, the vanes are simply supported, in spider form, as in an open design

As the impeller rotates, a quantity of gas is taken up into the passage between the vanes The shape of the impeller vanes and rotation speed move the gas along the passage, imparting energy to the gas Some of the energy is converted into the velocity of the gas Depending on the design, the balance of the energy conversion is to pressure As with all compressors,

a temperature rise is associated with the compression

Types. Centrifugal compressors are available in many different configurations, the most basic of which is the overhung single-stage compressor, that is, a single impeller in an overhung configuration (Fig 12) One method of compounding the centrifugal compressor is to use the overhung single-stage compressor with a gear Gas passes from one stage to another on the gear frame by way of coolers between the stages This is a common configuration for plant air, and is shown in cross section in Fig 13 An assembled gear-mounted centrifugal compressor is depicted in Fig 14 A form of multistaging is shown in Fig 15 Figure 16 shows a straight-through multistage centrifugal compressor built with a horizontally split casing

Fig 12 Single-stage compressor