COMPRESSOR HANDBOOK phần 9 ppsx

An anchored spring leaf inserted between the ring and journal raises the amount of oil delivery and extends the ring’s region of stable operation.. a frictionless guided valve plate usef

⬚ F

Pour point,

⬚ F

Approximate cost per

Nonflammable aircraft hydraulic oil.

Skydrole 3.85 15.5 ⬎ 20,000 355 70 12.00 Nonflammable hydraulic oil for diecasting machines,

punch pressures, etc.

⬚ F

Pour point,

⬚ F

Approximate cost per

Polyglycol:

LB-140X 5.7 29.8 — 345 ⫺ 50 2.40 Water-insoluble oils used for internal-combustion engines LB-300X 11.0 65.0 — 490 ⫺ 40 2.40 (Prestone Motor Oil), high temp bearings in ovens and

50-HB-55 2.4 8.9 — 260 ⫺ 85 2.40 Water-soluble oils used in wire drawing, metal forming,

Hydrolube 300N — 666.3 — None ⫺ 55 2.50 Water-polyglycol mixture used as non-flammable

hydraulic fluid in die-casting and machine tool work Chlorinated

aromatics:

Aroclor 1248 3.1 48 — 380 20 2.30 Die-casting machines and high-pressure compressors.

Polybutenes:

Fluorolubes:

Fluorolubes FS 1.10 3.52 — None — 300.00 Equipment handling liquid oxygen, concentrated

Fluorolubes S 4.6 24.1 — None — 300.00 hydrogen perioxide, etc Density of approximately 1.8

grams / cc.

Process and natural gas compressors.

15 10 9.0 8.0 7.0 6.0 5.0 4.0 3.0 20

SAE 20 W SAE 10

vy Steam Cylinder Oil

Gr ade 1010 Jet Engine OilLight Spindle Oil

Light Turbine and Electr

ic Motor Oil

Medium T urbine Oil

T, Fo

FIGURE 19.87 Viscosity of petroleum oils.

Reyns (lb forc-sec / m2) 6.895 ⫻ 10 3 Pascal-sec (N-sec / m 2 )

TABLE 19.29 General Types of Additives with Typical Chemical Compositions 36

Dithiophosphate

Sodium petroleum sulfonate

Boundary lubrication Chlorinated naphthalene

Sulfurized hydrocarbon Viscosity index improver Polyisobutylene Pour-point depressant Polymethnerylate

carbon-steel back; (2) an intermediate layer of copper or bronze; and (3) an overlay

of lead-base babbitt from 0.001 to 0.020 in thick The intermediate layers increasethe mechanical strength of the babbitt bearing and also provide reasonably goodbearing surfaces in cases the thin babbitt surface layer is destroyed in operation

Non-Babbitt Bearing Materials. Other common bearing materials used, ever babbitt cannot be employed are:

when-• Bronze Bearing bronzes may be grouped into lead bronzes, tin bronzes, andhigh-strength bronzes The strength and high-temperature properties generallyimprove as one proceeds from the high-lead to high-tin to various high-strengthbronzes However, there is a loss in the compatibility properties as the amount

of lead decreases For this reason, it is generally advisable to use the highest lead

66 ⬚ F 212 ⬚ F

Brinell hardness

68 ⬚ F 212 ⬚ F

Melting point

⬚ F

Complete liquefaction

Yield point*

psi

66 ⬚ F 212 ⬚ F

Ultimate strength*

psi

66 ⬚ F 212 ⬚ F

Brinell hardness

68 ⬚ F 212 ⬚ F

Melting point

⬚ F

Complete liquefaction

outstand-• Aluminum Aluminum bearing alloys offer excellent resistance to corrosion byacidic oils, good load-carrying capacity, superior fatigue resistance, and goodthermal conductivity A smooth machine finish of the running surface is recom-mended along with a clean lubricant, a shaft hardness of 300 Brinell or higher,and a large enough clearance to allow for the high thermal expansion of thealuminum Sometime the aluminum is overlaid with a thin coating of lead babbitt.This overlay assists in making up for the otherwise poor embeddability and con-formability characteristics of the aluminum.

The range of temperatures that these various bearing materials, as well as someother materials, can endure is given in Table 19.31

In practice, a designer must obtain quantitative data to ascertain on the one handwhether the bearing will meet his operational requirements, and on the other handfind out what the power losses, flows, temperatures, etc will be to properly planthe layout of the facility In Sections 19.3 to 19.7, the graphs and tables offer valuesfor the performance of various bearing designs These, however, do not exhaustthe information required for rational design What is needed is some orientationhow the various geometrical and operational parameters affect bearing operationand how to go about improving or even optimizing a given bearing design Thefollowing paragraphs should offer some guidance as to how to go about approach-ing this task

19.129

19.130 CHAPTER NINETEEN

The attitude angle ␾is defined as the angle between the line centers—a linepassing the centers of bearing and journal—and the load vector When the treat-ment is restricted to vertical loads,␾denotes the angle between location of hmin

and the vertical and therefore the importance of ␾lies in that it determines the

has traditionally been the most important parameter However, a more convenient

quantity is the inverse of S, here called the load parameter, given by

W ⫽ ␮N冉 冊R ⫽ LD␮N冉 冊R (19.59b) where P ⫽ ( W / LD) is the unit loading What this parameter says is that any combination of P, ␮, N, C, and R such as to leave the value of W unchanged,would result in the same bearing eccentricity ratio, ⑀, and attitude angle,␾

• Minimum film thickness The is the smallest distance between the journal andbearing surfaces and it is given by:

hmin

C What is normally referred to as load capacity relates to the load, W, which this

hmin can support

• Friction coefficient This is the ratio between the frictional force and bearingload It is normally expressed in the form of:

R (R/C)F␶

ƒ⫽

The general shape of ƒ as a function s is given in Fig 19.88 The region of

sudden rise in ƒ denotes the limit of hydrodynamic lubrication, followed by aregime of ‘‘boundary lubrication’’ characterized by partial contact between themating surfaces

• Power loss This, of course, can be obtained from the value of F␶, namely

H⫽ F␶R␻ ⫽ ƒWR␻

H⫽冉 冊H0 ⫽ [␲ ␮3 N LD /C]2 3 (19.61)

The quantity by which H is normalized, represents the power loss in an unloaded

concentric journal bearing, i.e., one in which ⑀ ⫽0 It is known as the Petroffequation

FIGURE 19.88 Behavior of friction coefficient in fluid film

bearings.

• Flow An amount of lubricant, Q1, enters the bearing at the leading edge; an

amount, Q s leaks out the two sides of the bearing (one-half Q sat each side), and

an amount Q2 leaves the trailing end of the pad In most cases, since a journalbearing extends over circumference (2␲), Q2 is not discharged outside but reen-ters the next oil groove, so that the net amount of lubricant to be made up from

an outside source is Q s The latter is referred to as side leakage Clearly we mustalways have

⑀ ⫽0 (for which case Q s⫽ 0 and Q1 ⫽ Q2)

The above flows, Q1, Q s , and Q2 are what may be called hydrodynamic flows

induced by the shearing action and pressure gradients of the fluid film Q sis theminimum amount of oil to be delivered to the bearing to maintain a full fluidfilm with all its potentialities In practice, designers supply more than this re-

quired minimum, using a supply pressure p s ⬎ p a The effect of the supplypressure, usually of the order of 10 to 30 psig, can be ignored as far as bearinghydrodynamics are concerned

• Temperature rise A bulk temperature rise can be estimated from the values ofpower loss and side leakage, namely

The coefficients⑀, ƒ,H, Q Q K1, 2, andBwhich serve to evaluate bearing ance are obtained from solutions of the Reynolds equation for the specific geom-etries and operating conditions of the various bearing designs Many such solutionswere given in Sections 19.3 through 19.7.

perform-19.9.2 Bearing Configuration

The behavior of a bearing is naturally a function of its geometry However, evenfor a given design there are a number of variables that will affect its performance

Among the more known parameters are the L /D and C /R ratios and the degree of

preload Of the less familiar ones one can cite load orientation, the geometry ofthe oil grooves or the relative proportions of a bearing’s geometrical elements

Journal Bearings. Although one often hears about the use of full, that is, 360⬚arc bearings, it is very rarely that such sleeves are employed in machinery Mostjournal bearings consist of two or more pads separated by horizontal oil groovesmaking them in fact partial bearings, used either singly or in tandem The numberand distribution of these angular pads on bearing performance is one of the moreimportant considerations in bearing design

Partial Bearings Whenever a single pad of an angular extent␤ ⬍2␲is used,

it is called a partial bearing When ␤ is very small, its load capacity is low, asillustrated in Figs 19.89 and 19.90 However, soon a limit is reached at about␤ ⫽

140⬚ beyond which no further gains are registered The reason for this asymptoticbehavior is due to oil cavitation at the trailing end of the pad where the pressuresdecrease close to or even below ambient pressure Thus, if a partial bearing is usedthere is no need to go beyond a 140⬚ arc The effect of temperature in partialbearings is a combination of two phenomena The higher the arc the longer thedissipation path and the higher the temperatures; however, a longer arc producesthicker films and thus less heating Consequently, as shown in Fig 19.91, a cross-over point occurs; at high loads low values of ␤ are preferred, if low ⌬T ’s are

desired; at low loads a longer arc is preferred

Grooved Bearings Partial bearings are not used extensively The most common

designs are grooved bearings which consist of a number of pads arranged in tandem

by cutting axial oil grooves around the 360⬚circumference There is a great variety

FIGURE 19.89 Effect of bearing arc on load capacity 21

of such designs, the most common being a 2-pad bearing with two grooves at thehorizontal split Others may have 3, 4 or 6 grooves forming the same number ofindividual pads The more grooves the lower the load capacity, as shown in Figs.19.92 and 19.93 Thus, if load capacity is the primary objective, a 2-groove bearing

is best; however, those with a larger number of grooves are somewhat more stable.Related to the above is the fact that any hole or disruption in the bearing surfacewill reduce the load capacity Figure 19.94 shows the effects on the pressure profile

of cutting a slit or circular hole in the loaded part of a bearing The larger theincursion, the more drastic the reduction in the hydrodynamic pressures whichtranslates directly into reduced load capacity

Tilting Pad Bearings The primary characteristic of this family of bearings is

that the individual pads are not fixed but are pivot-supported so that during ation not only does the journal move but so do the pads and each in a differentfashion A general picture of a tilting 3-pad bearing is shown in Fig 19.95 Thestructural and analytical complexities of these bearings are more than compensated

oper-19.134 CHAPTER NINETEEN

FIGURE 19.90 Effect of bearing arc on load capacity 21

by their great reliability and the fact that they have no rival in their stability acteristics

char-The number of possible design parameters and operating modes in a tilting padbearing is very large Some of them are discussed below

a Number of pads. Table 19.32 gives a comparison of a 3-pad versus a 5-padcentrally pivoted bearing having zero preload When the load is in line with thepivot, the 3-pad design has a higher load capacity but the reverse is true whenthe load direction is between the pads For loads of engineering interest the 5-pad design consumes less power

b Pivot location. In order to assure two-directional rotation and for ease of sembly, most tilting pad bearings are centrally pivoted However, a 10% or 15%displacement of the pivot in either direction would not significantly alter thegeneral performance, a slight preference being a downward shift

as-c Preload. From many standpoints a high preload is desirable Its effect on venting the scraping of the top pads has been discussed previously and from

pre-this standpoint an m of at least 0.5 is required High preloads also yield higher

stiffness and damping However, the penalty is that the film thickness over the

pivot and often also the absolute hminis reduced Likewise, the power losses andtemperatures rise with an increase in preload

d Mode of loading. In general, the shaft eccentricity will be lower when loadedover the pivot It is characteristic of tilting pad bearings that, regardless of

FIGURE 19.91 Effect of bearing arc on value of hmin 31

whether the load vector is over the pivot or between the pads, the locus of shaftcenter is along a vertical line which has a direct beneficial effect on stability.Results for the two modes of loading on stiffness and damping are given in Fig.19.96 for a bearing of zero preload As seen, both the spring and dampingcoefficients are lower for the between-pads mode of loading

Oil-Ring Bearings As pointed out previously, oil ring bearings operate under

starved conditions It is thus the main task of the designer to find ways to increase

as much as possible the amount of oil delivered to the bearing surface Some ofthe important parameters that play a role in accomplishing it are geometry shape

of contact surface, weight, the material and size of the ring relative to the shaft In

an experimental study, a series of rings portrayed in Table 19.33 was tested withthe purpose of both increasing the flow of lubricant and of extending the regime

of stable ring operation The conclusions reached were as follows:

a An optimum ring shape is one with a quasi-trapezoidal cross-section and a series

of straight teeth at the contact surface shown in Table 19.33 as Ring No 2

b The best ring material is bronze with a weight of 0.135 lb per inch of ring

circumference

c For bearing diameters in excess of 6 in., dual rings are recommended.

19.136 CHAPTER NINETEEN

FIGURE 19.92 Load capacity of grooved bearings 27

d An anchored spring leaf inserted between the ring and journal raises the amount

of oil delivery and extends the ring’s region of stable operation One such bilizer is shown in Fig 19.97

sta-Load Angle Bearing loads are usually directed midway of a pad or between

grooves However, improved performance can be obtained by shifting the loadvector toward the trailing edge of the bearing pad A comprehensive mapping ofthe effects of shifting the load vector around the circumference of a 2-groovebearing is shown in Fig 19.98 Normally the load would be straight down, that isalong␾L⫽0 However, as seen in the figure by moving the load toward the trailingedge, improved performance is obtained for the entire range of bearing operation

At low loads an optimum occurs at a load of␾L⫽ 10⬚; at high loads the value of

␾Lis some 30⬚ The lowest load capacity would occur at a load angle of 60⬚fromthe midway point Supplementary data is given in Table 19.34, where it is seenthat the worst angular position results in a load capacity reduction of 70 to 80%.Similar data for a 3-groove bearing has been given in Fig 19.92 Achievement of

an optimum bearing position requires no special effort It is sufficient to rotate thebearing in the housing the required 10⬚to 30⬚to obtain this Attention should only

be given to the oil delivery path since now the oil grooves would no longer be atthe horizontal split This can be taken care of by cutting a short oil supply channel

on the outside of the bearing shell

FIGURE 19.93 Comparisons of 2- and 4-axial groove

bearings.

Misalignment It was pointed out in an earlier section that an overhung impeller

will cause bearing misalignment As shown in Fig 19.99, in severe misalignmentthe journal at one end may find itself in the upper half of the bearing even thoughthe load is downward As a consequence, a fluid film and hydrodynamic pressuresmay develop in both the lower and upper portions of the bearing Stretching fromthe end where the hydrodynamic film is at the bottom, this film will wrap itself inhelical fashion around the entire bearing circumference In all cases the load ca-

pacity, that is the value of h for the imposed load, will be drastically reduced.

Thrust Bearings. Unit loads in thrust bearings are higher than in journal bearings

and consequently their hmin will be smaller But it should also be realized that,

except for a bearing with a flat at the end, hmin in thrust bearings occurs not along

a line as in journal bearings but at a point, namely the outer downstream edge of

the pad This point is also where Tmax will occur and again it will be higher than

in journal bearings This is due to the low value of hminbut also to the higher linearvelocities of the runner at the outer radius of the pad

FIGURE 19.94 Effect of a slot and a hole on hydrostatic pressure.

FIGURE 19.95 A tilting 3-pad journal bearing.

TABLE 19.32 Relative Load Capacity and 5-Pad Bearings 29

3-W

On-pivot load

3 pads 5 pads

Load between pivots

of (h1 ⫺ h2) and␤for the entire range of (L / R2) ratios From this an optimum set

of design parameters can be obtained for a particular application It is worth notingthat in general the optimum configuration is that which yields nearly square bearingpads

An improved version of a plain tapered land bearing is one with a flat surface

at the trailing end, as shown in Fig 19.36 The additional merit of this design isthat upon starting and stopping, the runner rides on a flat surface reducing wear

19.140 CHAPTER NINETEEN

FIGURE 19.96 Effect of mode of loading on bearing stability in a 4-pad tilting pad bearing.

Here a new parameter is the ratio of the tapered to the flat portion The plot in Fig.19.100 shows such a variation from 60% to 100% taper, the latter being the taperedland bearing discussed previously The load capacity peaks at a taper value of about80% of the pad arc, that is the tapered portion should be four times that of the flat.Interestingly the value of (power loss / load capacity) achieves a minimum at thesame point

Misalignment In properly operating thrust bearings, the load carried by each

pad is the same When the shaft and consequently the runner is misaligned, this is

no longer true and some of the pads are much more heavily loaded than the other

A pictorial representation of this situation is given in Fig 19.101 As seen, theloads carried by the heavily loaded pads as well as their maximum temperaturescan be 10 times as high as the ones located opposite them where the runner is

furthest from the pads The values of h in the two sets of pads will be of the same

ratio The span of severity of bearing operation goes up with the number of padsused in the misaligned bearing Thus, if misalignment is expected, one should notuse more than 4 to 6 pads

Hydrostatic Bearings In a conventional hydrostatic bearing portrayed in Fig.

19.102, the load capacity is given by

␲R ( p2 o⫺ p ) [1 a ⫺ (R /R ) ]1 2

W⫽

2 ln (R /R )2 1There are therefore two parameters that determine the level of W; ( p ⫺ p) and

TABLE 19.33 Oil Ring Configuration 15

(R2/R1) The variation of load with these quantities is shown in Fig 19.102 Asseen no optimum for load capacity occurs; it rises with ⌬p and drops with a rise

in (R1/ R2) A minimum occurs in the power loss, but power loss in a hydrostaticbearing is not of great concern and, when it is, it is due not to bearing geometrybut to the onset of turbulence in the fluid

of values listed in the table.

Nearly all the bearing data given here are for bearings operating under laminarconditions Should turbulence set in, the operating characteristics will change Onemay expect turbulence when the bearing Reynolds number reaches a value between

750 and 1500 The higher the Reynolds number, the more intense will be the effect

of turbulence Table 19.37 shows what will be the impact of the turbulent regime

on the major items of bearing operation

FIGURE 19.98 Effect of load angle on load capacity in two-groove bearing 37

TABLE 19.34 Effect of Load Angle on Load Capacity in

Conventional Two-Groove Bearing 13

Worst condition

19.144 CHAPTER NINETEEN

FIGURE 19.99 Hydrostatic forces and films under misalignment.

TABLE 19.35 Optimum Pad Arrangement 25

L / R2 h1/ ␦ ␪ ␤ , deg Number of pads

19.146 CHAPTER NINETEEN

FIGURE 19.101 Effects of misalignment in thrust bearings 28

FIGURE 19.102 Performance of incompressible

hydro-static bearing.

TABLE 19.36 Typical Design Limits for Journal Bearings

Minimum film thickness 0.001–0.01 in (0.0025 to 0.25 mm)

Temperature rise Up to 80 ⬚ F (27 ⬚ C) (on babbitt)

Maximum temperature Up to 300 ⬚ F (150 ⬚ C) (on babbitt)

80 ⬚ to 30 ⬚ for tilting pad

Inlet oil temperature, T 80 ⬚ F to 130 ⬚ F (27 ⬚ to 55 ⬚ C)

Code: ( ⫹ ) means increase magnitude of parameter to achieve effect in left-hand column.

( ⫺ ) means decrease magnitude of parameter to achieve effect in left-hand column.

Example: To decrease temperature rise, one or more of the following can be done:

decrease L / D ratio, increase C / R, decrease oil viscosity, etc.

Preload,

Supply oil pressure, ps

below

( ⫹ ) or ( ⫺ ) ( ⫹ ) ( ⫺ ) No effect

Pad

( ⫹ ) or ( ⫺ ) ( ⫹ ) ( ⫹ ) No effect

a) The stability of a journal bearing increases in the following order: circular, pressure, elliptical,

3-lobe, tilting pad.

b) Apparent effect only.

TABLE 19.39 Characteristics of Various Journal Bearings Journal Bearing Summary

Table

Axial Groove 1 Easy to make

2 Low cost

1 Subject to oil whirl Round bearings are nearly

always ‘‘crushed’’ to make elliptical or multi-lobe Elliptical 1 Easy to make

2 Low cost

3 Good damping at critical speeds

1 Subject to oil whirl at high speeds

2 Load direction must be known

Probably most widely used bearing at low or moderate speeds

Three and

Four Lobe

(Tapered

Land, etc.)

1 Good suppression of whirl

2 Overall good performance

3 Moderate cost

1 Some types can be expensive

to make properly

2 Subject to whirl at high speeds

Currently used by some manufacturers as standard bearing design

3 Load direction must be known

Very popular with petrochemical industry Easy to convert elliptical over to pressure dam.

Hydrostatic 1 Good suppression of oil whirl

2 Wide range of design parameters

3 Moderate cost

1 Poor damping at critical speeds

2 Requires careful design

3 Requires high pressure lubricant supply

Generally high stiffness properties used for high precision rotors Tilting Pad 1 Will not cause whirl (no cross

coupling)

2 Wide range of design parameters

1 High cost

2 Requires careful design

3 Poor damping at critical speeds

4 Hard to determine actual clearances

5 High horsepower loss

Widely used bearing to stabilize machines with subsynchronous non-bearing excitations

19.10 REFERENCES

The selection of the following references was made with the intent of providingsources from which additional data could be culled for bearing design purposes

1 Allaire, P E., D F Li, and K C Choy, ‘‘Transient Unbalance Response of Four

Mul-tilobe Journal Bearings,’’ Journal of Lubr Technology, Trans ASME, July 1980.

2 Chen, H M., ‘‘Active Magnetic Bearing Technology: A Conventional Rotordynamic Approach,’’ 15th Leeds-Lyon Symposium on Tribology, September 1988.

3 Chen, H M., ‘‘Magnetic Bearings and Flexible Rotor Dynamics,’’ STLE Annual Meeting

at Cleveland, Ohio, May 9–12, 1988.

4 Chen, H M., et al., ‘‘Stability Analysis for Rotors Supported by Active Magnetic ings,’’ 2nd International Symposium on Magnet Bearings, July 12–14, 1990, Tokyo, Japan, pp 325–328.

Bear-5 Chen, H M., ‘‘Design and Analysis of a Sensorless Magnetic Damper,’’ presented at ASME Turbo Expo, June 5–8, 1995, Houston, Texas, 95GT180.

6 Compressor Handbook, Gulf Publishing Co., Book Division.

7 Gross, W A., ‘‘Gas Film Lubrication,’’ John Wiley, 1962.

8 Heshmat, H., J A Walowit, and O Pinkus, ‘‘Analysis of Gas-Lubricated Compliant Thrust Bearings,’’ ASME Paper 82-LUB-39, 1982.

9 Heshmat, H., J A Walowit, and O Pinkus, ‘‘Analysis of Gas-Lubricated Foil Journal Bearings,’’ ASME Paper 82-LUB-40, 1982.

10 Heshmat, H., and J Dill, ‘‘Fundamental Issue in Cryogenic Hydrodynamic Lubrication,’’ Proc AFOSR / ML Fundamentals of Tribology Workshop (February 1987).

11 Heshmat, H., ‘‘Analysis of Compliant Foil Bearings with Spatially Variable Stiffness’’ presented at AIAA / SAE / ASME / ASEE 27th Joint Propulsion Conference, June 24–26,

1991, Sacramento, CA, Paper No AIAA-91-2101.

12 Heshmat, H., ‘‘A Feasibility Study on the Use of Foil Bearings in Cryogenic pumps,’’ presented at AIAA / SAE / ASME / ASEE 27th Joint Propulsion Conference, June 24–26, 1991, Sacramento, CA, Paper No AIAA-91-2103.

Turbo-13 Heshmat, H., and P Hermel, ‘‘Compliant Foil Bearing Technology and Their Application

to High Speed Turbomachinery,’’ 19th Leeds-Lyon Symposium on Thin Film in ogy—From Micro Meters to Nano Meters, Leeds, U.K., Sept 1993, D Dowson, et al (eds) (Elsevier Science Publishers B.V., 1993), pp 559–575.

Tribol-14 Heshmat, H., and O Pinkus, ‘‘Performance of Starved Journal Bearings with Oil Ring

Lubrication,’’ Journal of Tribology, Trans ASME 107, no 1 (1985): 23–32.

15 Heshmat, H., and O Pinkus, ‘‘Experimental Study of Stable High-Speed Oil Rings,’’

Journal of Tribology, ASME 107, no 1 (1985): 14–22.

16 Heshmat, H., and O Pinkus, ‘‘Performance of Oil Ring Bearing,’’ International Science Conf on Friction, Wear, Lubr., Tashkent, U.S.S.R., May 1985.

17 Hustek, J F., and O J Peer, ‘‘Design Considerations for Compressors with Magnetic Bearings,’’ Proc 3rd Int Symposium on Magnetic Bearings, July 1993, Alexandria, VA.

18 Jones, G J., and F A Martin, ‘‘Geometry Effects in Tilting-Pad Journal Bearings,’’ ASLE Paper No 78-AM-@A-2, 1978.

19.152 CHAPTER NINETEEN

19 Ku, C.-P R., and H Heshmat, ‘‘Compliant Foil Bearing Structural Stiffness Analysis:

Part I—Theoretical Model Including Strip and Variable Bump Foil Geometry,’’ Journal

of Tribology, Trans ASME, vol 114, no 2 (1992): 394–400.

20 Pinckney, F D., and J M Keesee, ‘‘Magnetic Bearing Design and Control Optimization for a Four-Stage Centrifugal Compressor,’’ Proceedings of Mag ’92, pp 218–227.

21 Pinkus, O., and B Sternlicht, ‘‘Theory of Hydrodynamic Lubrication’’ (New York: McGraw-Hill, 1961).

22 Pinkus, O., and D F Wilcock, ‘‘Low Power Loss Bearings for Electric Utilities: Volume II: Conceptual Design and Optimization of High Stability Journal Bearings; Volume III: Performance Tables and Design Guidelines for Thrust and Journal Bearings,’’ MTI Re- port Nos 82TR42, 82TR43, April 1982.

23 Pinkus, O., ‘‘Analysis of Elliptical Bearings,’’ Trans ASME, vol 78, 1956, pp 965– 973.

24 Pinkus, O., ‘‘Analysis and Characteristics of the Three-Lobe Bearing,’’ Trans ASME, Ser.D., vol 81, March 1959.

25 Pinkus, O., ‘‘Solution of the Tapered-Land Sector Thrust Bearing,’’ Trans ASME, vol.

80, Oct 1958.

26 Pinkus, O., ‘‘Analysis of Non-circular Gas Journal Bearings,’’ Journal of Lubr

Technol-ogy, Trans ASME, Oct 1975.

27 Pinkus, O., ‘‘Solution of Reynolds Equation for Arbitrarily Loaded Journal Bearings,’’ Trans ASME, Series D, vol 83, no 2, June 1961.

28 Pinkus, O., ‘‘Misalignment in Thrust Bearings Including Temperature and Cavitation

Effects,’’ Journal of Tribology, Oct 1986.

29 Pinkus, O., ‘‘Optimization of Tilting Pad Journal Bearings Including Turbulence and

Thermal Effects,’’ Israel Journal of Technology, vol 22, nos 2–3, 1984 / 85.

30 Pinkus, O., ‘‘Manual of Bearing Failure and Repair in Power Plant Rotating Equipment,’’ EPRI, July 1991.

31 Raimondi, A A., and J Boyd, ‘‘A Solution for the Finite Journal Bearing and Its plication to Analysis and Design—III,’’ Trans ASLE, vol l, no l, 1959.

Ap-32 Reddickoff, J M., and J H Vohr, ‘‘Hydrostatic Bearings for Cryogenic Rocket Engine

Turbopumps,’’ Journal Lubr Technology, July 1969.

33 Schmied, J L and J C Predetto, ‘‘Rotor Dynamic Behaviour of a High-Speed Oil-Free Motor Compressor with a Rigid Coupling Supported on Four Radial Magnetic Bear- ings,’’ Proceedings of 4th International Symposium on Magnetic Bearings, August 23–

26, 1994, ETH Zurich, Switzerland, pp 441–447.

34 Vohr, J H., ‘‘The Design of Hydrostatic Bearings,’’ Columbia University, NY.

35 Walton, J F., and H Heshmat, ‘‘Compliant Foil Bearings for Use in Cryogenic pumps,’’ Proceedings of Advanced Earth-to-Orbit Propulsion Technology Conference Held at NASA / MSFC May 17–19, 1994, NASA CP3282, vol 1, Sept 19, 1994, pp 372–381.

Turbo-36 Wilcock, D F., and Booser, E R., ‘‘Bearing Design and Application’’ (New York: McGraw Hill, 1957).

Consulting Mechanical Engineer

Hoerbiger Corporation of America, Inc.

a frictionless guided valve plate useful for non-lubricated compressors, a feature,which until the introduction of non metallic materials to compressor valves waspoorly achieved by most other valve designs

In 1910, Hans Mayer patented a different valve design that became known asthe ‘‘Feather valve.’’ This design uses several flexible steel strips as sealing ele-ments and became the standard for the Worthington Company for many years

In 1931, Ingersoll Rand patented yet another valve design that became known

as the ‘‘Channel valve.’’ The channel valve employed several strips with a crosssection like a ‘‘U’’ (therefore the name ‘‘channel’’), each supported by a leaf spring

It was the standard valve for the Ingersoll Rand Company and was probably duced in higher numbers than any other compressor valve

pro-A valve design older than all the others was rediscovered in the late 1950’s withthe event of pipeline compressor applications and the availability of high strengthplastics, in this case nylon Thompson Industries developed this poppet valve for

20.2 CHAPTER TWENTY

the Clark Brothers, a compressor manufacturer in Olean, New York This valveproved to be extremely efficient in low compression ratio applications

All automatic compressor valves have several basic components in common:

• Valve seat

• Sealing element(s)

• Lift constraint (guard)

• Spring(s)

The flow passages in a valve can be arranged in various different ways:

• Circular rows of ports

• Parallel series of ports

• Irregular arranged number of holes

20.3.1 Valve Designs Used in Air and Gas Compressors

Ring Valve. Ring valves are probably the most commonly used valves in air aswell as gas compressors A properly manufactured valve ring is a very simpleelement with a perfectly uniform stress distribution It therefore has a high tolerancefor impacts From a valve designer’s point of view, rings can easily be guided andthe utilization of the available area is good (for manufacturing reasons most valvesare round) The disadvantage of individual rings is the need for each ring to beseparately spring loaded Since the specific spring load for each ring cannot beperfectly identical and because the flow distribution across a valve is not uniform,

it is difficult if not impossible to make all rings move uniformly On the other hand,due to the possibility of individual motion of each ring, this design is more tolerant

to liquids than other designs

Several ring valve designs which are used are discussed below

Simple Ring Valve This design uses plain rings guided in the valve guard The

valve springs can either be small individual coil springs, separate for each ring, orslightly larger springs supporting two rings Some older valve designs use one largecoil per ring Wafer or lentoid springs are used in smaller valves The advantage

of these springs is their low space requirement, allowing for extremely thin valveguards and therefore low clearance volume The disadvantage of these springs isthe limited possibility of adaptation to different operating conditions For this rea-son, they are mostly used in air compressors

Damped Ring Valve The only damping system used in ring valves is gas ing This design uses very thick valve rings guided on the full diameter in a closely

damp-fit groove in the valve guard In theory, when the valve opens, the gas beneath the

FIGURE 20.1 Ring and spring set for ring valve with one spring covering 2 rings.

FIGURE 20.2 Ring valve with wafer springs.

valve ring is trapped in the guard groove and has to be squeezed out through thenarrow passages on either side of the valve ring The effect of damping largelydepends on production tolerances, state of wear, and / or presence of liquid or solidcontamination in the gas On valves with non-metallic valve rings and steel guards,the different coefficient of thermal expansion makes this an almost impossible taskfor the designer It follows that the application for this damping system is limited

to metallic sealing elements

Contoured Ring Valves Some valve designs utilize a contoured valve ring

to-gether with a heavily chamfered or contoured valve seat groove This is done toachieve a lower flow resistance in the lift area where, with conventional designs,the gas has to make two 90⬚turns On the other hand, for given valve lift and giventotal length of seat lands, the geometric valve port area is reduced in the proportion

of sine of the angle of flow deviation

All these valves use non-metallic valve ring materials Since these materials have

a different thermal expansion coefficient than the valve seat material, there may be

a leakage problem either in the cold or hot condition of the valve This problem isreduced by a certain flexibility of the valve ring which allows it to ‘‘roll into’’ thevalve seat groove

20.4 CHAPTER TWENTY

DAMPENING BY TRAPPED GAS

VALVE SEAT

VALVE GUARD

VALVE PLATE CONTROLLED

GAP

FIGURE 20.3 Gas dampened ring valve.

FIGURE 20.4 Contoured ring

valve (courtesy Cook Manly).

Ported Plate Valve. There are basically two types of ported plate valves in ance:

exist-Simple Plate Valve Plate valves, for the most part, are ring valves with their

individual rings connected by bridges The advantage of a plate valve over anindividual ring valve is that there is only one sealing element to be controlled Thesimultaneous opening or closing of all ports is automatically given This advantagehowever comes at a price From a stress point of view, the simple, very uniformvalve ring has been transformed into a much more complicated element When avalve plate impacts at one point it starts to bend, causing a rather complex, nonuniform stress distribution

Double or Mass Damped Plate Valve The designation double damped valves

is actually a misnomer, since there is only one real damping feature used Thevalve springs themselves are not to be considered a damping feature

Plate valves allow for a very effective, mechanical damping system So calleddamping plates are positioned between the valve plate and the valve guard andsometimes are spring loaded separately During the valve opening event, the valve

FIGURE 20.5 Double damped valve.

FIGURE 20.6 Velocity diagram of mass damped valve.

plate travels the first portion of the lift alone and then collides with the dampingplates During this impact, linear momentum is conserved, but kinetic energy isdestroyed In other words, the energy required to accelerate the mass of the damp-ing plates results in a reduction of the velocity of the valve plate itself and con-sequently reduces the final impact velocity against the valve guard As an extrabonus, the valve plate tends to be leveled when contacting the damping plates

Channel Valve. By numbers, this valve design is by far the most widely usedcompressor valve in the Western hemisphere It uses a number of straight sealing

elements with the U shaped cross section, therefore the name channel valve Each

individual channel is spring loaded by a leaf spring and guided on both ends of

20.6 CHAPTER TWENTY

FIGURE 20.7 Channel valve (courtesy of

Dresser Rand ).

FIGURE 20.8 Feather valve

(courtesy of Dresser Rand ).

the channel by a comb like guide This valve design is very efficient in low tomedium pressure, low to medium speed compressors

Feather Valve. This valve design is intriguing for its apparent simplicity It uses

a leaf type sealing element which is allowed to bow into a machined recess in theguard The leaf is therefore also its own spring Feather valves are no longer used

in new compressors and their application was limited to low and medium pressurecompressors with clean gas service

Poppet Valve. Poppet valves use rather large (approx 7 / 8 inch diameter) holes

in the valve seat and for each of these holes there is a mushroom shaped sealingelement, called a poppet Each poppet has its own valve spring The original pop-pets were made of bronze, which due to its high mass, was practically useless Theadvent of Nylon made this design the valve of choice for low compression ratio,low speed compressors Valve lifts of up to 3 / 8 inch were commonly used and inpipeline service this valve’s efficiency was unequaled

FIGURE 20.9 Poppet valve (courtesy

of Dresser Rand ).

When PEEK (Polyetheretherketone,1a polymer) became available, poppet valveswere also successfully used in higher compression ratios With reduced valve liftsnecessary for these applications, the efficiency of the standard poppet valve should

be checked against other valve designs

For high speed compressors, a variation of the common poppet valve was

de-veloped This design utilizes a much smaller poppet—called Minipoppet—and

therefore a smaller valve seat hole Lower valve lifts without reducing flow areaare possible with this design

Double Deck Valve. In conventional cylinder designs, there is limited space able for compressor valves If the valve area achieved is insufficient and addedclearance volume can be tolerated, double deck valves can be used Double deckvalves can be of different basic valve designs such as ring-, plate-, poppet- orfeather valves In all but the feather valve design, a second valve is positionedupside down above the first valve The two valves are separated by a spacer andheld together by means of a sleeve The gas flow to the bottom deck is straightfrom the cylinder; the flow to the top deck is through the sleeve and then reversed

avail-by means of a cover The combined valve area of a double deck valve is imately 40% larger than that of the same size single deck valve

approx-Deck-and-One-Half Valve. Deck-and-one-half valves are valves where a full deckvalve is positioned above a ring type bottom valve and the flow through both valves

is in the same direction These valves provide an even larger valve area than doubledeck valves, but require a very deep valve pocket since the cylinder porting has to

be above the top deck valve A cone type spacer is normally used to separate thetwo valve decks and provides flow access to the top deck The total valve area of

a deck-and-one-half valve can be 60 to 65% larger then the same size single deckvalve

in refrigeration compressors and small air compressors Small displacement highpressure cylinders also use this valve design.

Reed Valve. The domain of the reed valve encompasses small refrigeration andair compressors A reed valve normally consists of a single seat plate with valvereeds positioned on both sides of the seat plate The reeds themselves are made ofthin strip steel and can have almost any shape The reed is held on one end andcovers a port with the other end The valve opens when the differential pressurestarts to bend the reed away from the seat plate This elastic deformation of thereed also acts as the valve spring and no other springs are used

There are reed valve designs using two seat plates with a spacer (normally agasket) and the valve reeds positioned between the seat plates The advantage ofthis design is a positive stop for the valve reed when the valve is open; the dis-advantage is its high clearance volume and a higher cost compared to a single seatplate design Due to the low mass of the moving element—the valve reed—thisdesign can be used at very high compressor speeds Reed valves are commonlyused in compressors up to 3600 rpm and have been successfully tested at speeds

up to 7000 rpm