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October 2009 19882009 Chevron U.S.A. Inc. All rights reserved. 200-1

200 Centrifugal Compressors

Abstract

This section discusses engineering principles, types of machines and configurations,

and performance characteristics. It contains sufficient information, when used in

conjunction with Company specifications, to understand how to specify and apply

centrifugal compressors including auxiliaries and support systems.

The discussion is primarily aimed at heavy-duty multistage units, but the

information can be applied to smaller and less severe-duty compressors as well.

Contents Page

210 Engineering Principles 200-3

211 Gas Flow Path

212 Conversion of Velocity Energy to Pressure

213 Thermodynamic Relationships

214 Performance Related to Component Geometry

215 Compressor Types

220 Performance Characteristics 200-13

221 General

222 Impeller Performance Curves

223 Use of Fan Laws

224 Surge

225 Stonewall

230 Selection Criteria 200-25

231 Application Range

232 Horsepower and Efficiency Estimates

233 Head/Stage

234 Stages/Casing

235 Discharge Temperature

236 Selection Review

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200 Centrifugal Compressors Compressor Manual

200-2 19882009 Chevron U.S.A. Inc. All rights reserved. October 2009

240 Machine Components and Configurations 200-32

241 Machine Components

242 Dry Gas Seals

243 Configurations

250 Application and System Considerations 200-80

251 Effect of System Changes on Performance

252 Stable Operating Speed Ranges

253 Power Margins

254 Series Operation

255 Weather Protection

256 Process Piping Arrangements

257 Lube- And Seal-Oil Systems

260 Instrumentation and Control 200-88

261 Typical Instrumentation

262 Compressor Control

263 Control System Selection

264 Surge Control

265 Machinery Monitoring

270 Rerates and Retrofits 200-92

271 Capacity

272 Pressure

273 Power

274 Speed

280 Foundations 200-95

281 Foundation Mounting

282 Design Basis for Rotating Compressors

290 Materials 200-99

291 Sulfide Stress Cracking

292 Stress Corrosion Cracking293 Hydrogen Embrittlement

294 Low Temperature

295 Impellers

296 Non-Metallic Seals

297 Coatings

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Compressor Manual 200 Centrifugal Compressors


210 Engineering Principles

This section covers the fundamentals of centrifugal compressors, describing the gas

flow path, conversion of velocity to pressure, thermodynamic relationships, and the

effect of component geometry on compressor performance.

These fundamentals provide a foundation for troubleshooting performance prob-lems, making rerating or initial selection estimates, evaluating vendor proposals,

engineering compressor applications, and assisting with overall process design.

211 Gas Flow Path

A discussion of the flow path through the centrifugal compressor will provide a

better understanding of the compression process.

There is often confusion about the term stage when applied to centrifugal

compressors. The process designer thinks of a stage as a compression step made up

of an uncooled section, usually consisting of several impeller/diffuser units. The

mechanical engineer or machine designer defines a stage as one impeller/diffuserset, and a section as a single compressor casing containing several stages. In this

section of the manual:

Stage isdefined as one impeller/diffuser set

Process stageis defined as an uncooled section (or casing) containing several

impellers/diffusers

Based on this, a centrifugal compressor is made up of one or more stages; each

stage consisting of a rotating component or impeller, and the stationary components

which guide the flow into and out-of the impeller. Figure 200-1shows the flow path

through a section of a typical multistage unit.

Fig. 200-1 Compressor Sect ion(Courtesy of the Elliot t Company)

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212 Conversion of Velocity Energy to Pressure

Pressure is increased by transferring energy to the gas, accelerating it through the

impeller. Note that all work on the gas is done by the impeller; the stationary

components only convert the energy added by the impeller. Part of this energy is

converted to pressure in the impeller and the remainder is converted to pressure as it

decelerates in the diffuser. A typical pressure-velocity profile across a stage is

shown in Figure 200-2.

Since the kinetic energy is a function of the square of the velocity, the head(not

pressure) produced is proportional to the square of the impeller tip speed:

(Eq. 200-1)where:

Note Head is a term often used for the work input to a compression process.

The units of head are foot-pounds (force) divided by pounds (mass) [newton meter

divided by kilograms]. In general practice, head is usually taken as feet or

meters.

Fig. 200-2 Pressure and Velocity Profile

U.S. Units Metric Units

H = head, ftlbf

lbm

N/m

kg

U = impeller tip speed in ft/sec m/sec

K = a constant a constant

g = 32.174 (ft-lb: mass) / (lb: force) (sec2) 1 m kg / N sec2

H KU2

g-------=

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where:

As mentioned in Section 100, the polytropic process is typically used for centrifugal

compressors (rather than the adiabatic process).

Using the relationship for k, n, andp, polytropic efficiency is:

(Eq. 200-4)

214 Performance Related to Component GeometryEffects resulting from the geometric shape of the principle components of the

compressor are shown in Figure 200-4. Variables such as the impeller configuration

and blade angle, inlet guide vane angle, diffuser size and shape, etc., can be adjusted

by the machine designer for optimum performance under a specified set of

operating conditions. Figure 200-5shows impeller vector diagrams for various

blade angles.

ZavgZ1 Z2+

2-------------------=

average compressibility=

p

k 1

k------------

n 1

n------------

------------=

Fig. 200-4 Impeller Inlet and Outlet Flow Vector Triangles (From Compressors: Selection & Sizing,by Royce Brown

1986 by Gulf Publishing Company, Houston, TX. Used with permission. All rights reserved.)
http://cmp-en-100.pdf/http://cmp-en-100.pdf/

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Impellers with backward leaning blades, are more commonly used for most

centrifugal compressors because of their increased stable operating range

(Figure 200-6). Forward and radial blades are seldom used in petrochemicalapplications.

Fig. 200-5 Forward, Radial, and Backward Curved Blades (From Compressors: Selection & Sizing, by Royce Brown

1986 by Gulf Publ ishing Company, Houston, TX. Used with permission. All rights reserved.)

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Machine output is always affected by combined losses, such as:

Mechanical loss

Aerodynamic loss

Friction and shock loss

Mechanical losses, such as those from a journal or thrust bearing, affect the power

input required, but do not influence the head-capacity curve. Aerodynamic losses

that do influence the shape of the curve consist mainly of wall friction, fluid shear,

seal losses, recirculation in flow passages, and shock losses. Shock lossesare the

result of expansion, contraction, and change of direction associated with flowseparation, eddies, and turbulence. Friction and shock lossesare the predominant

sources of the total aerodynamic losses.

Fig. 200-6 Effect of Blade Angle on Stability

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Figure 200-7illustrates the affect of these combined losses in reducing the

theoretical head.

Friction losses can be reduced by improving surface finishes. Shock losses may

sometimes be mitigated by further streamlining of flow passages. These techniques

will improve efficiency and tend to reduce the surge point, but they are costly, and

there is a point of diminishing returns. The Company specification does not allow

the manufacturers quoted performance to include efficiency improvements due to

impeller polishing.

215 Compressor Types

There are two types of compressors, defined by either an axial or radial casing

construction. Figure 200-8illustrates this construction, referred in the API 617

Standards as:

axial, or horizontally split

radial, or vertically split

Fig. 200-7 Typical Compressor Head

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API 617 (Centrifugal Compressors) requires the use of the vertically-split casings

when the partial pressure of hydrogen exceeds 200 psi (13.8 bar).

Other factors which influence the horizontal/vertical split decision include theabsolute operating pressure of the service and ease of maintenance for a particular

plant layout.

The top half of the horizontally-split casing (Figure 200-9) is removed to access the

internals. The stationary diaphragms are installed individually in the top and bottom

half of the casing. Main process connections may be located either in the top or

bottom half.

The horizontally-splitdown-connected casing has the advantage of allowing

removal of the top half for access to the rotor without requiring removal of major

process piping.

Vertically-splitor barrel compressors have a complete cylindrical outer casing. Thestationary diaphragms are assembled around the rotor to make up an inner casing,

and installed inside the outer casing as a unit, contained by heads or end closures at

each end. Some later designs hold the heads in place by use of shear rings

(Figure 200-10).

Fig. 200-8 Joint Construct ion (Courtesy of the Howell Training Group)

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On the vertically-split casing, maintenance of the rotor and other internal parts

(other than bearings and shaft-end seals) involves removal of at least one head,

withdrawal of the inner casing from the outer pressure containing casing, and then

dismantling of the inner casing to expose the rotor (Figure 200-11). The inner casing

and rotor can be removed from either the up- or down-connected vertically-splitouter casing without disturbing process piping.

Both the horizontally and vertically-split casing designs allow removal of bearings

and shaft-end seals for maintenance without disassembly of major casing

components.

Fig. 200-9 Horizontally-split Casing(Courtesy of the Howell Training Group)

Fig. 200-10 Shear Ring Head Retainer (Courtesy of Dresser-Rand)

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Figure 200-12gives a comparison of pressure vs. capacity for multistagehorizontally- and vertically-split casing construction. The size/rating comparisons

are general. Specific pressure/capacity ranges and casing configurations vary

between manufacturers.

Fig. 200-11 Vertically-split Casing (Courtesy of the Howell Training Group)

Fig. 200-12 Pressure/Capacity Chart (Courtesy of Dresser-Rand)

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Overhung-Impeller Types

Single-stage, overhung-impeller (impeller located outboard of the radial bearings,

opposite the driver end) designs are available in pressure ratings to approximately

2000 psi (138 bar) and capacities to 50,000 cfm (85000 m3/hr).

Another type of centrifugal compressor is the integrally-geared configuration. Thisis an overhung-impeller type built around a gear box, with the impellers attached to

gear pinion shafts and the impeller housings mounted on the gear box. Possible

configurations include two, three, four, and even five stage designs with capacities

to 30,000 cfm (51,000 m3/hr) and pressures to 250 psig (17 bar). These have

typically been used as packaged-air or nitrogen compressors. The overall

arrangement of this type varies significantly between manufacturers.

Major features of the integrally geared design include:

Open impellersmaximum head developed

volute diffusers for optimum efficiency

different pinion speeds to optimize impeller efficiency

220 Performance Characteristics

221 General

Figure 200-13presents a centrifugal compressor performance map, using API 617

nomenclature. The family of curves depicts the performance at various speeds

where N represents RPM, and:

Vertical axisHead:polytropic head, pressure ratio, discharge pressure, or

differential pressure; and

Horizontal axisInlet Capacity: called Q or Q1 shown as actual inlet

volume per unit of time ACFM or ICFM where A is actual, or I is inlet.

Note that inlet flow volume, or capacity, is based on a gas with a particular

molecular weight, specific heat ratio, and compressibility factor at suction pressure

and temperature.

The curve on the left represents the surge limit. Operation to the left of this line is

unstable and usually harmful to the machine.

A capacity limitor overload curveis shown on the other side of the map. The area

to the right of this line is commonly known as stonewallor choke. Operation

in this area is, in most instances, harmless mechanically, but the head-producing

capability of the machine falls off rapidly, and performance is unpredictable.

Surge and stonewall should not be confused. Although machine performance is

seriously impaired in either case, they are entirely different phenomena. These are

covered in more detail later in this section.

Terms frequently used to define performance are stability range and percent

stability. Referring again to Figure 200-13, the rated stability range is taken as

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QD- QSwhere QDis the rated point and QSis the surge point along the 100 percent

speed line. The percent stability expressed as a percentage is:

(Eq. 200-5)

% stabilityQD QS

QD

--------------------- 100=

Fig. 200-13 Typical Centrifugal Compressor Performance Map (Courtesy of the American Petroleum Insti tute)

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222 Impeller Performance Curves

For convenience, manufacturers usually base the performance of individual

impellers on an air test. Figure 200-14represents a typical curve which

characterizes a certain impeller design. The vertical axis is usually called the head

coefficient ; and the horizontal axis is called the flow coefficient,. (See

Section 212for definitions of and ). In this way, impeller performance data are

concisely cataloged and stored for use by designers. When a compressor is

originally sized, the designer translates the wheel curve data into ACFM, discharge

pressure, and RPM in wheel-by-wheel calculations to select a set of wheels that

satisfy the purchasers requirements.

Theoretically, an impeller should produce the same head, or feet of the fluid,

regardless of the gas weight. However, in practice, a wheel will produce somewhat

more head (than theoretical) with heavy gases, and less with lighter gases. Gas

compressibility, specific heat ratio, aerodynamic losses, and several other factors are

responsible for this deviation. Manufacturers should apply proprietary correction

factors when the effect is significant. This effect contributes to variance from the

well-known fan lawsor affinity laws. (See the next sub-section.)

Notice in Figure 200-14that the heavier gas causes surge at a higher Q/N, that is, it

reduces stability. The opposite is true of a lighter gas. Similar non-conformance can

sometimes be observed when the wheel is run at tip speeds considerably higher orlower than an average design speed. The higher tip speed would surge at higher

Q/N, and the lower tip speed would surge at a lower Q/N.

Figure 200-15illustrates the effects of using movable inlet guide vanes. Notice that as

the head or discharge pressure is reduced, the surge volume (defined by the dashed

line) is also reduced. The effect is similar to that of speed reduction on a variable

speed machine. Inlet throttling, although less efficient, will produce similar curves.

Fig. 200-14 Individual Impeller Performance Curve

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Centrifugal compressors recognize actual inlet cubic feet per minute (ACFM at inlet

conditions, or ICFM). Performance curves are most commonly plotted using

ACFM. This means that a curve is drawn for a specific set of suctionconditions,

and any change in these conditions will affect the validity of the curve.

Performance curves often plot discharge pressure on the vertical axis, and flow

(ACFM) on the horizontal axis. To estimate performance for varying suction

pressures, the curve should be converted to pressure ratioon the vertical axis. This

can be done by dividing the discharge pressures on the vertical axis by the suction

pressure on which the original curve was based. The effect of a small variation insuction temperature can be estimated by using a ratio of absolute temperatures with

the original temperature in the denominator. This ratio is used to correct the inlet

capacity on the X-axis by multiplying inlet capacities by the temperature ratio.

For a rough estimate for molecular weight changes of less than 10 percent, the

pressure ratio on the curve can simply be multiplied by the ratio of the new

molecular weight over the original. Unless there are gross changes in the gas

composition causing large changes in specific heat ratio, this estimating method will

only have an error of 12 percent for pressure ratios between 1.5 and 3. For more

accurate estimates, a curve with polytropic head on the vertical axis must be

obtained.

Remember that any change that increases the density of the gas at the inlet willincrease the discharge pressure and the horsepower. Also, the unit will tend to surge

at a slightly higher inlet volume.

Fig. 200-15 Constant Speed Machine with Variable Inlet Guide Vanes

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223 Use of Fan Laws

Fan lawscan be used in many cases to estimate performance for small changes in

speed and flow, but care and judgment must be used. Using these laws is risky, and

should be done cautiously.

The fan laws state that inlet volume is proportional to speed, and that head isproportional to the speed squared. These laws are based on the assumption that the

fluid is non-compressible. Fan laws may be inaccurate when testing the

performance level of multistage compressors at off-design speeds. (Figure 200-16

illustrates this error.) Similar errors could be incurred in estimating surge volumes

using the fan laws.

To illustrate, assume a 10 percent mass flow reduction to the first stage. If all other

inlet conditions remain the same, volume flow will also be reduced by 10 percent.

Since mass flow was reduced by 10 percent, the second stage will also see a

10 percent flow reduction. (Figure 200-13shows that flow reduction results in an

increased discharge pressure from the first stage.) Since volume is inversely

proportional to pressure, the volume to the second stage will be reduced further inproportion to the increased discharge pressure from the first stage. The second stage

will have a similar effect on the third stage and so on. Deviation from the ideal gas

laws will increase significantly as the number of compressor stages increases.

224 Surge

Surgeis a situation that can destroy a compressor. It is a critical factor in design of

the compressor and its control system. It is also a critical operating limit.

Surge is a condition of unstable flow within the compressor, resulting in flow

reversal and pressure fluctuations in the system. This occurs when the head

(pressure) developed by the compressor is less than that required to overcomedownstream system pressure. At surge, continuous forward flow is interrupted.

While surge is caused by aerodynamic instability in the compressor, interaction with

the system sometimes produces violent swings in flow, accompanied by pressure

fluctuations and relatively rapid temperature increase at the compressor inlet. Surge

affects the overall system and is not confined to only the compressor. Therefore, an

understanding of both the external causes and the machine design is necessary to

apply an adequate anti-surge system.

The compressor surge region was previously identified in Figure 200-13. In

Figure 200-17lines depicting three typical system operating curves have been

added. The shapes of these curves are governed by the system friction, and pressure

control in the particular system external to the compressor

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200-18

19882009ChevronU.S.A.

Inc.Allrightsreserved.

October2009

Fig. 200-16 Error in Fan Laws Multistage Compressor

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A compressor will operate at the intersection of its curve and the system curve.To change the point at which the compressor operates:

1. Change the speed or variable geometry of the compressor, thus relocating the

compressor curve; or

2. Change the system curve by repositioning a control valve or otherwise altering

the external system curve.

Typical Surge Cycle

A typical surge cycle is represented by the circuit between points B, C, D, and back

to B (Figure 200-17). If events take place which alter the system curve to establish

operation at point B, the pressure in the system will equal the output pressure of the

compressor. Any transient can then cause reverse flow if the compressor dischargepressure falls below the downstream system pressure.

For reverse flow to occur, compressor throughput must be reduced to zero at point C

which corresponds to a pressure called the shut-off head. When the system

pressure has decreased to the compressors shut-off head at C, the machine will re-

establish forward flow since the flow requirement of the compressor is satisfied by

the backflow gas (compressor capability now greater than system requirements).

Fig. 200-17 Typical Centrifugal Compressor Performance Map Showing Surge Cycle

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Now that the compressor has sufficient gas to compress, operation will immediately

shift to the right in approximately a horizontal path to point D. With the compressor

now delivering flow in the forward direction, pressure will build in the system, and

operation will follow the characteristic speed curve back to points B and C. The

cycle will rapidly repeat itself unless the cause of the surge is corrected, or other

favorable action taken, such as increasing the speed.

Several internal factors combine to develop the surge condition. From the surge

description, you can see that the domed shape of the head-capacity characteristic

curve is fundamentally responsible for the location of the surge point at a given

speed. On the right side of the performance map (Figure 200-17) the slope of the

curve is negative. As inlet flow is reduced, the slope becomes less negative until it

reaches zero at the surge point. As flow is reduced further to the left of the surge

point, the slope becomes increasingly positive.

Section 210covers internal factors and their effect on location of the surge region.

Frequency of Surge

Frequency of the surge cycle varies inversely with the volume of the system. For

example, if the piping contains a check valve located near the compressor discharge

nozzle, the frequency will be correspondingly much higher than that of the system

without a check valve. The frequency can be as low as a few cycles per minute up to

15 or more cycles per second. Generally, the higher the frequency, the lower the

intensity. The intensity or violence of surge tends to increase with increased gas

density which is directly related to higher molecular weights and pressures, and

lower temperatures. Higher differential pressure generally increases the intensity.

Design Factors Affecting Surge

A greater number of impellers in a given casing will tend to reduce the stable range.

Similarly, so does the number of sections of compression, or the number of casingsin series.

The large majority of centrifugals use vaneless diffusers, which are simple flow

channels with parallel walls, without elements inside to guide the flow. The

trajectory of a particle through a vaneless diffuser is a spiral of about one-half the

circumferential distance around the diffuser (Figure 200-18). If this distance

becomes longer for any reason, the flow is exposed to more wall friction which

dissipates the kinetic energy. As flow is reduced, the angle is reduced which extends

the length of the trajectory through the diffuser (Figure 200-19). When the flow path

is too long, insufficient pressure rise (head) is developed and surge occurs.

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Fig. 200-18 Design Condition Velocity Triangles(Reproduced with permission of the Turbomachinery Laboratory. From

Proceedings of the Twelfth Turbomachinery Symposium, Texas A&M Universi ty, College Station, TX,1983)

Fig. 200-19 Flow Trajectory in a Vaneless Diffuser(Reproduced with permission of the Turbomachinery Laboratory.

From Proceedings of the Twelfth Turbomachinery Symposium, Texas A&M University, College Station, TX,1983)

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Occasionally, vaned diffusers are used to force the flow to take a shorter, more effi-

cient path. Figure 200-20shows the flow pattern in a vaned diffuser. The vaned dif-

fuser can increase the aerodynamic efficiency of a stage by approximately 3 percent,

but this efficiency gain results in a narrower operating span on the head-capacity

curve with respect to both surge and stonewall. The figure also shows how the path

of a particle of gas is affected by off-design flows. At flows higher than design,impingement occurs on the trailing side of the diffuser vane creating shock losses

which tend to bring on stonewall. Conversely, flow less than design encourages

surge, due to the shock losses from impingement on the leading edge of the vane.

Despite adverse effects on surge, the vaned diffuser should be applied where

efficiency is of utmost importance, particularly with small high-speed wheels.

Stationary guide vanes may be used to direct the flow to the eye of the impeller.

Depending upon the head requirements of an individual stage, these vanes may

direct the flow in the same direction as the rotation or tip speed of the wheel, an

action known as pre-rotationor pre-swirl. The opposite action is known as

counter-rotationor counter swirl. Guide vanes set at zero degrees of swirl are

called radial guide vanes.

The effect guide vanes have on a compressors curve is illustrated in Figure 200-21.

Note that pre-rotation reduces the head or unloads the impeller. Pre-rotation tends to

reduce the surge flow. Counter-rotation increases the head and tends to increase the

surge flow.

Fig. 200-20 Vaned Diffuser

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Movable inlet guide vanes are occasionally employed on single-stage machines, or

on the first stage of multistage compressors driven by electric motors at constant

speed. The guide vane angle can be manually or automatically adjusted while the

unit is on stream to accommodate operating requirements. Because of the

complexity of the adjusting mechanism, the variable feature can only be applied to

the first wheel in almost all designs.

External Causes and Effects of Surge

Briefly, some of the usual causes of surge (other than from machine design) are:

1. Restricted suction or discharge such as a plugged strainer.

2. Process changes in pressures or gas composition.

3. Mis-positioned rotor or internal plugging of flow passages.

4. Inadvertent speed change such as from a governor failure.

The effects of surge can range from a simple lack of performance to serious damage

to the machine and/or the system. Internal damage to labyrinths, diaphragms, thrust

bearing and the rotor can be experienced. Surge often excites lateral shaft vibration.It can also produce torsional damages to such items as couplings and gears.

Externally, devastating piping vibration can occur causing structural damage,

mis-alignment, and failure of fittings and instruments.

Surge can often be recognized by check valve hammering, piping vibration, noise,

wriggling of pressure gages or ammeter on the driver. Mild cases of surge are

sometimes difficult to discern.

Fig. 200-21 Effect of Guide Vane Setting (Stationary or Variable)

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225 Stonewal l

Another major factor affecting the theoretical head-capacity curve is chokeor

stonewall. The terms surge and stonewall are sometimes incorrectly used

interchangeably, probably due to the fact that serious performance deterioration is

observed in either case.

A compressor stage is considered to be in stonewall, in theory, when the Mach

number equals one. At this point the impeller passage is choked and no more flow

can be passed. Industry practice normally limits the inlet Mach number to less than

0.90 for any specified operating point.

We are concerned with two important items in defining stonewall: the inlet-gas

velocity incidence angle, and the inlet-gas Mach number.

The vector diagram (Figure 200-22) shows an inlet-gas velocity vector which lines

up well with the impeller blade at design flow.

The ratio of the inlet gas velocity (relative to the impeller blade) to the speed of

sound at inlet is referred to as the relative inlet Mach number.

(Eq. 200-6)

where:

As flow continues to increase, the incidence angle of the relative gas velocity, with

respect to the impeller blade, becomes negative as shown in Figure 200-23. The

negative incidence angle results in an effective reduction of the flow area and

impingement of the gas on the trailing edge of the blade, contributing to flow

separation and the onset of choke.

Fig. 200-22 Inlet Gas Velocity Vector Design Flow(Courtesy of the Elliot t Company)

Mach No.Vre l

a1----------=

a1 g k ZRT1=

speed of sound at inlet=

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It is important to note the choke effect is much greater for high molecular weight

gas, especially at low temperatures and lower k values. For this reason, maximum

allowable compressor speed may be limited on high molecular weight applications,

with a corresponding reduction in head per stage.

230 Selection Criter ia

This section concentrates on equipment selection. (Forms are also available in the

Appendix to assist in the estimating process.)

231 Application Range

Refer to Figure 200-12for a chart of capacity vs. pressure for horizontally- and

vertically-split centrifugal compressors.

Normally, manufacturers do not design a compressor to match an application, theyfit the application to one of a series of existing compressor casings or frame sizes.

Therefore, check the manufacturers bulletins for data required to make selection

estimates. Figure 200-24provides data for a series of compressor casings based on a

comparison of data from the industry.

Fig. 200-23 Inlet Gas Velocity Vector Negative Incidence Angle (Onset of Choke)

(Courtesy of the Elliott Company)

Fig. 200-24 Preliminary Selection Values for Multistage Centrifugal Compressors (1 of 2)

Casing Size

(Frame)

Flow Range

ft3/min (m3/hr)

Head Coefficient

(1)Nominal Impeller

Diameter (inches/mm)

Typical Speed

RPM

Maximum(2)

Polytropic Head/Stage

ft (m)

1 5007,000

(85011,900)

0.48 1416

(335.6406.4)

12,000 12,000

(35,880)

2 2,00018,000

(3,40030,600)

0.490.50 1822

(457.2558.8)

9,000 12,000

(35,880)

3 6,00031,000

(10,20052,700)

0.500.51 2430

(609.6762)

6,500 12,000

(35,880)

4 10,00044,000

(17,00074,800)

0.500.51 36

(914.4)

5,600 12,000

(35,880)

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In addition, the minimum discharge volume flow should be considered. Current

impeller designs limit impeller inlet flow to approximately 300500 cfm

(500850 m3/hr). Thus, process conditions resulting in actual discharge volume of

less than approximately 250 cfm (425 m3/hr) may be unacceptable.

232 Horsepower and Efficiency Estimates

One of the major benefits in doing your own estimates, rather than turning

everything over to a manufacturer, is that you develop a better understanding of the

application. You are then in a better position to discuss it with the manufacturers,

evaluate alternate selections, and even catch errors in manufacturers estimates.

Figure 200-25is a plot of polytropic efficiency vs. inlet volume flow. This chart

may be used for estimating polytropic efficiencies.

As discussed in Section 100, manufacturers use a computer to calculate compressor

performance on a stage-by-stage basis. Performance is based on each preceding

stage, new impeller inlet conditions, including compressibility (Z) and k values todetermine the individual performance for each successive stage.

If specific stage data is unavailable, overall calculations using average

compressibility and a k value based on the average flange-to-flange temperature,

will provide reasonably accurate results. (Refer to Section 100for compressibility

equations.)

Estimate overall efficiency from Figure 200-25, using average CFM from:

(Eq. 200-7)

where discharge ACFM is determined using Equation 200-14and an efficiency

of 75 percent.

5 30,00065,000

(51,000

110,000)

0.510.52 4245

(1,066.81,143)

4,400 12,000

(35,880)

6 60,000100,000

(102,000170,000)

0.520.53 54

(1,371.6)

3,600 10,000

(3,048)

7 80,000160,000

(136,000

272,000)

0.530.54 60

(1524)

3,000 10,000

(3,048)

(1) Based on backward leaning blades

(2) For 2830 molecular weight at normal temperatures

Fig. 200-24 Preliminary Selection Values for Multistage Centrifugal Compressors (2 of 2)

cfmavgInlet ACFM Disch. ACFM+

2---------------------------------------------------------------------=
http://cmp-en-100.pdf/http://cmp-en-100.pdf/http://cmp-en-100.pdf/http://cmp-en-100.pdf/

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Determine n-1/n from:

(Eq. 200-8)

Recalculate head, discharge temperature, and gas horsepower (GHP) from:

(Eq. 200-9)

where:

Hp = Polytropic Head in foot-lbf/lbm (N-m/kg)

(Eq. 200-10)

Fig. 200-25 Polytropic Efficiency vs. Inlet Volume Flow (Courtesy of Dresser-Rand)

n 1

n------------

k 1

kp------------=

Hp zavg RT1r

n 1

n------------

1

n 1

n------------

---------------------=

T2 T1r

n 1

n------------

=

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(Eq. 200-11)

where:

w = weight flow in lbs/min (kg/sec)

Estimate shaft power using:

Shaft power = Gas power + bearing loss + seal loss

where bearing lossis determined from Figure 200-26, and oil seal loss is

determined from Figure 200-27. The casing size in the figures is selected by

comparing the cfmavgwith the flow range in Figure 200-24.

Gas Power HPwHp

33 000 p,-------------------------=

Gas Power kWwHp

1 000 p------------------------=

Fig. 200-26 Bearing Losses vs. Casing Size and Speed(Courtesy of Dresser-Rand)

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233 Head/Stage

Although special impeller designs are available for higher heads, a good estimate

for the typical multistage compressor is approximately 10,000 ft/stage

(3048 m/stage). This is based on an assumed impeller flow coefficient of 0.5 and a

nominal impeller tip speed of 800 fps (244 mps).

The actual head per stage varies between manufacturers and individual impeller

designs, ranging from 9,00012,000 feet (27433658 meters) for 2830 molecular

weight gas at normal temperatures.

Head per stage is limited by:

impeller stress levels

inlet Mach number

Impeller Stress Level

The following speed margins are defined by API:

Fig. 200-27 Oil Seal Losses vs. Casing Size and Speed (Courtesy of Dresser-Rand)

Rated (Design) Speed: 100%

Maximum Continuous Speed: 105% of Rated Speed

Trip Speed: 110% of Maximum Continuous

Overspeed: 115% of Maximum Continuous

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Figure 200-28identifies the impeller stresses at various rotational speeds. Reduced

yield strengths required for corrosive gas will correspondingly reduce maximum

head per stage through reduction in speed.

Inlet Mach Number

An increase in gas molecular weight, or a decrease in k, Z or inlet temperature will

result in an increase in inlet Mach number. For high molecular weight or low

temperature applications, Mach number may limit head per stage for a given design.

234 Stages/Casing

The maximum number of stages per casing should normally be limited to eight. It is

usually limited by rotor critical speeds, although in a few cases temperature can be a

limiting factor.

Most multistage centrifugal compressors operate between the first and second

criticals (flexible shaft rotor). Figure 200-29shows the location of critical speeds in

relation to the operating speed range. API specifies the required separation between

critical speeds and the compressor operating range. As the bearing span is increased

to accommodate additional impellers, the critical speed decreases, with the second

critical approaching the operating range. While some manufacturers bulletins

indicate as many as 10 or more stages per casing, designs exceeding eight impellers

per case should be carefully evaluated against operating experience from similar

units.

For compound, or sidestream loads, additional stage spacing may be required to

allow for intermediate exit and/or entry of the gas. In these applications, the numberof impellers would be reduced accordingly.

Fig. 200-28 Impeller Stress Levels at Various Speeds

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235 Discharge TemperatureIf the calculated discharge temperature exceeds approximately 350F (177C),

cooling should be considered to avoid problems with compressor materials, seal

components, and clearances. The exact temperature limit is dependent on factors

such as the gas compressed, compressor materials, allowable temperature of the seal

oil, and the type of seals. Also, note that discharge temperature will increase as flow

is reduced toward surge.

236 Selection Review

Refer to Section 2100 for centrifugal compressor checklists, which provide typical

items covered during the review of any centrifugal compressor quotation.

Fig. 200-29 Rotor Response Plot(Courtesy of the American Petroleum Institute)

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240 Machine Components and Configurations

241 Machine Components

Centrifugal compressors are made up of a casing with stationary internals,

containing a rotating element, or rotor, supported by bearings. Shaft end-seals are

provided to contain the process gas. Figure 200-30shows a typical multistage

compressor and identifies the basic components. (Refer to Figure 200-1for details

of the gas flow path.

Casings

Nozzles

Stage

Diaphragms

Impellers

Rotor

Shaft Radial Bearings

Thrust Bearing

Balance Piston

Interstage Seals

Shaft-end Seals

The main machine components are:

Casings

The following is a summary of casing materials and their applications.

1. Cast Iron

Limited to low pressure applications for non-flammable, non-toxic gases.

Limited in location and size of main and sidestream connections to avail-

able patterns.

2. Cast Steel

Quality is difficult to obtain.

X-ray inspection requirements increase costs.

High-rejection rate or involved repairs can extend deliveries.

3. Fabricated Steel

Used for both horizontally- and vertically-split casings.

Improved quality control possible.

Delays associated with rejection or repair of castings are avoided.

Variable stage spacing provides minimum bearing span for required

stages.)

Main and sidestream nozzle size and location are not limited by pattern

availability.

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October2009

19882009Chevro

nU.S.A.Inc.Allrightsreserved.

200-33

Fig. 200-30 Centrifugal Compressor Nomenclature (Courtesy of Demag Delaval)

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4. Forged Steel

Used for small vertically-split casing sizes where application involves very

high pressures.

All centrifugal compressor casings used to be cast. But, due to the problems

associated with quality control on large castings, coupled with improved fabricationtechniques and costs, many manufacturers converted to fabricated steel casings,

especially on the larger frame sizes.

Nozzles

Inlet and outlet nozzles are available in a variety of configurations, depending on

the manufacturer. They are normally flanged. (Typical arrangements are shown later

in this section.) API 617 covers requirements for flange type, and ratings of main

and auxiliary connections.

The increased use of fabricated cases has provided additional flexibility in nozzle

orientation.

If the installation permits, the following should be considered:

1. Horizontally-split units with process connections in the lower half (down-

connected) allow removal of the top half, and internals including rotor, without

disturbing the process piping.

2. If overhead process piping is required, the use of vertically-split barrel

compressor casings still allow removal of the inner casing and access to the

internals without removing process piping. Fabricated casing design makes the

vertically-split unit a cost-effective alternative for larger medium pressure

applications.

Stage

The heart of the centrifugal compressor is the impeller stage. The stage is made

up of the following parts (illustrated in Figure 200-31):

inlet guide vanes

impeller

diffuser

return bend (crossover)

return channel

The stage can be separated into two major elements:

The impellers which are mounted on the shaft as part of the rotor.

The stationary components including the inlet nozzle and other components

mentioned above.

The inlet volute, or return channel, guides the gas to the eye of the impeller, and

aided by the guide vanes, distributes the flow around the circumference of the

impeller eye.

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One method of adjusting the stage performance, is to use different guide vane

angles. This changes the angle of incidence on the impeller which in turn varies the

head, efficiency, and stability. There are three types of fixed guide vanes; radial,

against-rotation, and with-rotation. The influence of various guide vane angles on a

given impeller head characteristic is shown in Figure 200-32.

Diaphragms

The stationary members inside the casing are called diaphragms. The diaphragm

includes a diffuser for the gas as it leaves the impeller, and a channel to redirect thegas through the return bend and return channel into the next stage. Diaphragms can

be either cast or fabricated, with cast diaphragms normally made of iron. Normally,

diaphragms are not exposed to high pressure-differentials, and therefore are not

highly stressed. Diaphragms should be made of steel where high-differentials may

exist (such as back-to-back impellers).

Impellers

The impeller is the most highly stressed component in the compressor. Available

types vary widely, although the three basic types are designated as open, semi-open

and closed:

Openimpellers have the vanes positioned in a radial direction and have noenclosing covers on either the front or back sides.

Semi-openimpellers usually have the vanes positioned in a radial or backward

leaning direction and have a cover on the back side which extends to the periphery

of the vanes. The radial blade, semi-open impeller provides for a maximum amount

of flow and head in a single stage, even in large diameter impellers (Figure 200-33).

Fig. 200-31 Centrifugal Compressor Stage Components(Courtesy of the Elliott Company)

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Closedimpellers have enclosing covers on both the front and back side. This is the

most common type in our large process compressors. The blades are usually

backward leaning, although they may be radial. Forward leaning blades arenormally used only in fans or blowers. (See Figure 200-33.)

Single-inletimpellers take the gas in an axial direction, on one side of the impeller

only, and discharge the gas in a radial direction.

Fig. 200-32 Head-Capacity Characteristics of Constant Speed Centrifugal Compressor with

Capacity Regulated by Variable Inlet Vane Angle (Courtesy of Dresser-Rand)

Fig. 200-33 Impeller Types Closed and Semi-Open Backward Leaning(Courtesy of Dresser-

Rand)

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Double-flowimpellers take the gas in an axial direction, on both sides of the

impeller, and discharge the gas in a radial direction. They are, in effect, the

equivalent of two single-inlet impellers placed back-to-back and, in general will

handle twice the flow at the same head as a single-inlet impeller of the same

diameter operating at the same speed.

Some impeller designs utilize a three-dimensional blade or vane configuration,

which varies the inlet blade angle from hub to outside diameter, thereby providing

optimum aerodynamic geometry, and improved performance over that of two-

dimensional designs.

Centrifugal compressor impellers discharge gas radially, but the gas enters in an

axial direction. An axial flow element called an inducer is sometimes incorporated

into the impeller. This combination is called a mixed-flow impeller. This

configuration results in increased efficiency in high-flow applications.

In the past, riveted impeller construction was used in a large number of applications.

Today, construction with welded components is more common.

Rotor

The rotor is made up of the shaft, impellers, impeller spacers, thrust collar, and the

balance drum. Figure 200-34shows several rotor configurations with various

impeller types.

If a rotor always operates below the lowest critical speed, it is known as a stiff-shaft

rotor. In contrast, a rotor with a normal operating range above one or more of its

criticals is a flexible-shaftrotor. Most multistage centrifugal compressors have

flexible-shaft rotors; and therefore, must pass through at least one critical during

start-up or shutdown. From an operational point of view, stiff shafts would be

preferable. However, it is not practical since the shafts would become prohibitively

large.

Shafts

Shafts are made from alloy steel forgings, finished by grinding or honing to produce

the required finish. Special requirements are detailed in API 617 for balancing and

concentricity during rotor assembly. Impellers are normally mounted on the shaft

with a shrink fit with or without a key, depending on the particular manufacturer and

compressor frame size. Most manufacturers use shaft sleeves to both locate

impellers and provide protection for the shaft in the event of contact with internal

labyrinth seals.

Special attention must be given to minimizing mechanical and electrical runout at

the shaft area observed by proximity probes. See the General Machinery Manualformore information on mechanical/electrical mount.

Radial Bearings

Radial bearings on centrifugal compressors are usually pressure lubricated. For ease

of maintenance, they are horizontally-split with replaceable liners or pads. The

liners or pads are usually steel backed with a thin lining of babbitt.

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Fig. 200-34 Centrifugal Compressor Rotor Configurations (Courtesy of the Elliott Company)

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Since centrifugal rotors are relatively light, bearing loads are low. This often leads

to instability problems which must be compensated for by the bearing design. Due

to instability, the straight-sleeve bearing is used only in some slow-speed units with

relatively short bearing spans. The pressure-damsleeve bearing, and the tilting-

padbearing are two commonly used designs which improve rotor stability.

The top half of the pressure-dam design is relieved as shown in Figure 200-35,

creating a pressure point where the dam ends. This conversion of oil-velocity into

pressure adds to rotor stability by increasing the bearing load.

Fig. 200-35 Pressure Dam Sleeve Bearing Liner (Courtesy of the Elliott Company)

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The tilting-pad bearing shown in Figure 200-36is usually made up of five

individual pads, each pivoted at its midpoint. By adjustments to the shape of the

pads and bearing clearance, bearing stiffness and damping characteristics can be

controlled. This bearing is successful in applications where the pressure-dam design

is inadequate.

Thrust Bearing

The tilting padis the most common thrust bearing used in centrifugal compressors.

The flat landand tapered landbearings are used less frequently. Figure 200-37

shows a tilting-pad bearing, consisting of a thrust collar (collar disk) attached to the

rotor shaft, and a carrier ring which holds the pads. A button on the back of the pad

allows the pad to pivot freely, thus allowing adjustment to varying oil velocity at

different compressor speeds. A further refinement to the basic design is the self-

equalizing bearing shown in Figure 200-38. An equalizing bar design allows the

bars to rock until all pads carry an equal load.

Fig. 200-36 Tilting-Pad or Pivoted Shoe Radial Journal Bearing (Courtesy of the Elliott Company)

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Fig. 200-37 Button-Type Tilting-Pad Thrust Bearing (Courtesy of the Elliott Company)

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Balance Piston

Figure 200-39represents the pressure profile acting on a centrifugal compressor

impeller, showing net pressure and net thrust pattern. This pressure pattern on the

impeller results in a net thrust force towards the suction end of the machine. The

total net thrust is the sum of the thrusts from all the individual impellers.

Fig. 200-38 Self-Equalizing Tilting-Pad Thrust Bearing (Courtesy of the Elliott Company)

Fig. 200-39 Impeller Pressure and Thrust Patterns (Courtesy of the Elliott Company)

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The rotors thrust is handled by the thrust bearing. However, in most multistage

compressors, a very large, if not impractical, thrust bearing would be required to

handle the total thrust load, if not otherwise compensated. Therefore a thrust

compensating device, or balance piston(or balancing drum) is normally provided

as part of the rotating element.

As shown in Figure 200-40, compressor discharge pressure acts on the inside end of

the balance piston. The area on the discharge side (outside) is vented, usually to

suction pressure. The resulting differential pressure across the balance piston

develops a force which opposesthe normal thrust force, thus greatly reducing the

net thrust transmitted to the thrust bearing.

Thrust compensation can be regulated by controlling the balance piston diameter.

However, there are usually physical and design limitations. Normally a balancing

force less than the total impeller thrust (approximately 75 percent) is selected to

maintain the rotor on one face of the thrust bearing for all operating conditions.

Otherwise, the rotor could bounce back and forth between the thrust faces as

process conditions vary.

Interstage Seals

Internal seals are installed on multistage centrifugals to prevent leakage between

stages, thereby improving performance. Labyrinth seals are commonly used, being

located at the impeller eye and at the shaft between stages. Figure 200-41illustrates

internal labyrinth seals.

Fig. 200-40 Centrifugal Compressor Balance Drum

(Balance Piston) (Courtesy of the Howell

Training Group)

Fig. 200-41 Interstage Seals (Courtesy of Dresser-Rand)

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Shaft End-Seals

Centrifugal compressors use shaft end-seals to:

1. Restrict or prevent leakage of air or oil vapors into the process gas stream.

2. Restrict or prevent leakage of process gas from inside the compressor.

Various types of seals are used, depending on the gas being compressed, the pressures

involved, safety, operating experience, power savings, and process requirements.

Shaft end-seals are separated into two broad categories:

the restrictive sealwhich restricts but does not completely prevent leakage;and

the positive sealdesigned to prevent leakage.

Restrictiveseals are usually labyrinths. They are generally limited to applications

involving non-toxic, non-corrosive, abrasive-free gases at low pressures. In some

cases, ports for injection or withdrawal of the gas are used to extend the range of

effectiveness. Some possible arrangements are shown in Figure 200-42.

Another form of the restrictive seal is the dry carbon ringseal, often used on

overhung single-stage compressors where maximum sealing and minimum axial

shaft spacing are important. Since this seal can be held to close clearances, leakage

is less than with the labyrinth seal. Also, less axial shaft space is required (see

Figure 200-43).

Positiveseals, while varying somewhat in design between manufacturers, are either

liquid-filmor mechanical contacttype.

The liquid-film type is shown in Figure 200-44. A schematic of a seal system is

shown in Figure 200-45. Sealing oilis fed to the seal from an overhead tank located

at an elevation above the compressor set to maintain a fixed five psi (typically)

differential above seal reference pressure. (Seal reference pressure is very close tosuction pressure.)

The oil enters between the seal rings and flows in both directions to prevent inward

leakage to the process gas or outward leakage of the gas to the atmosphere. Buffer

ports are often available for injection of an inert gas to further ensure separation of

the process from the sealing medium. The oil-film seal is suitable for sealing

pressures in excess of 3000 psi (207 bar). (See Figure 200-46for an illustration of a

buffer-gas injection.)

The tilting-padoil seal (shown in Figure 200-47) is a design that recognizes that in

some cases the seal operates as a bearing. It can be used in high-pressure, high-

pressure-rise applications to improve rotor stability.

The mechanical contactseal (Figure 200-48) is used at pressures up to 1000 psi

(70 bar), and has the added feature of providing more positive sealing during

shutdown. Sealing is provided by means of a floating carbon ring seal riding

between a stationary and a rotating face. The seal medium (oil) functions primarily

as a coolant. Seal oil differential is controlled by a regulator rather than an overhead

tank.

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Fig. 200-42 Ported Labyrinth Seals(Courtesy of

the Elliott Company)

Fig. 200-43 Buffered Dry Carbon-Ring Seal (Courtesy of

the Elliott Company)

Fig. 200-44 Liquid (Oil) Film Seal (Courtesy of Dresser-

Rand)

Fig. 200-45 Oil Film Seal Schematic(Courtesy of

Dresser-Rand)

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Fig. 200-46 Oil Film Seal with Buffer to Separate Seal Oil from Bearing Oil (Courtesy of Dresser-Rand)

Fig. 200-47 Tilt-Pad Oil Film Seal (Courtesy of Dresser-Rand)

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Fig. 200-48 Mechanical Contact Seal (Courtesy of the Elliott Company)

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242 Dry Gas Seals

Dry gas seals represent the latest technology for compressor shaft end sealing, and

are currently the preferred sealing technology for most centrifugal compressor

applications. Under dynamic (rotating) conditions, dry gas seals function as

restrictive seals. Depending on the design and conditions, dry gas seals can behave

either as restrictive or positive seals under static conditions. Similar to pump

mechanical seals, dry gas seals use mating rings (faces) as the sealing interface

between the rotating and stationary parts. The seals depend on a fine balance

between pressure forces, closure spring forces, and aerodynamic forces that are

created by very shallow grooves or depressions typically on the rotating seal face, as

shown in Figure 200-49(second cross-section). This balance results in a face gap of

about 0.00010.0002 inches (35m), through which the seal leaks at very low

rates. Leak rates are generally dependent on seal size, sealing pressure, and

rotational speed, and are influenced to a lesser extent on gas conditions. Depending

on these parameters, leakage rates generally range from fractional SCFM to about

4 SCFM. Although the dry gas seal design concept first achieved significant

commercial use in the early 1980s, it can be traced back to the early 1950s. Drygas seal technology is presently also applied in pumps, and to a lesser degree, steam

turbines, but this section addresses only centrifugal compressor applications. Dry

gas seals are an advancing technology in the petrochemical industry, so it is

important to be aware of the age of information (including this section), as well as

the duration of successful field experience for any given design advancement.

Chevron Engineering Standard CMP-SU-5.02 is a purchase specification for dry gas

seals and related support systems, and should be used for new compressor and

retrofit applications in both upstream and downstream facilities. CMP-SC-5.02 is a

commented version of the specification.

Fig. 200-49 Dry Gas Seal Rotating Face Segment, Shown with Exaggerated Depth Groove Geometry.

(Second cro ss-sectional view sho ws op erating face gap.)Courtesy of Flowserve Corporation

Face Rotation

GasPath

Rotating Face

Stationary

Face

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In general, dry gas seals offer the following primary advantages compared to other

sealing technologies:

significantly lower leakage rates and far greater pressure capability vs. other

restrictive type seals (labyrinth seals), and

simpler, more efficient and lower cost operation and auxiliaries vs. other

positive type seals (oil film seals, mechanical seals).

Dry gas seals can offer additional advantages as well, all of which should be

considered in the economics if justification for gas seals is needed (see Application

Considerations). Justification is usually an issue for retrofits, but on new

compressors, economics are favorable, especially if the alternative design requires

expensive and/or inefficient auxiliaries (seal oil systems, eductor systems, etc.).

The primary advantages of dry gas seals are the result of an advanced and precise

design that relies heavily on the proper operating environment. Reliable operation is

extremely dependent on having an adequate and uninterrupted supply of seal gas

(the gas the seal faces are exposed to) that is free of particulates and liquids. In

addition, the reliability of these seals can be compromised when the designapproaches current experience envelopes in sealing pressure, sealing temperature,

and seal face surface speeds, either singularly or in combination. All of these issues

focus on assuring the proper gas film and stress levels at the seal faces. Other

vulnerabilities may include seal face hang-up (which alters the seal face gap),

reverse rotation, reverse pressurization, and lube oil contamination of the seal faces.

These vulnerabilities are mitigated to a large degree through design of the dry gas

seal as well as the supporting auxiliary systems.

Arrangements

Depending on the application, one or two pairs of faces may be used in three basic

arrangements, usually in conjunction with labyrinth seals, to achieve the desired

process gas containment level. One pair of faces (a single seal) may be used for

moderate pressure applications that are not flammable, toxic nor environmentally

harmful (air, nitrogen), since the normal seal leakage will be to atmosphere.

However, low pressure services suitable for a single seal are also suitable for

labyrinth seals, which offer greater simplicity and reliability, as well as significantly

lower initial cost. A single seal arrangement is shown in Figure 200-50.

More typical applications require a dual seal arrangement to further limit or prevent

leakage to atmosphere, and provide a higher level of containment integrity. Dual

seals can be provided in either a double sealarrangement or a tandem seal

arrangement. Double seals are oriented in an opposed fashion to contain seal gas

(also called barrier gasin double seals) supplied between the inner and outer seals

(see Figure 200-51).

The barrier gas must be available at all times at a pressure higher than the process

gas pressure at the seals (or the sealing pressure). The sealing pressure is usually

very close to suction pressure during operation, but a compressor trip can cause

sealing pressure to rise to a settle-out pressurein some compressor circuits.

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The double arrangement is generally desirable only when nitrogen can be used as

the barrier gas since it provides a reliable, consistent and easily treatable supply of

seal gas, and also assures complete process gas containment with very low nitrogen

consumption. The double arrangement is also desirable for very low sealingpressures when there is a high potential for primary seal reverse pressurization of a

tandem arrangement (see Seal Gas Supply Objectives). The double arrangement

allows a small amount of barrier gas leakage both into the compressor across the

inner seal, and also to atmosphere across the outer seal.

Fig. 200-50 Simplified Single Seal Arrangement, shown without Primary Seal Labyrinth(Courtesy of Flowserve

Corporation)

Fig. 200-51 Simplified Double Seal Arrangement shown without Primary Seal Labyrinth (Courtesy of Flowserve

Corporation)

Clean Seal Gas

PROCESS ATMOSPHERE

Leakage

PROCESS ATMOSPHERE

Seal Gas

(Barrier Gas)

Leakage Leakage

Inner Seal Outer Seal

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Properly filtered nitrogen provides both dry and clean conditions for both inner and

outer seals, assures zero process gas emissions to atmosphere and requires a

relatively simple auxiliary system. Since a very small amount of leakage is inward

toward the compressor, it is necessary to verify that the process is tolerant of small

amounts of nitrogen.

In most services, especially where the process gas is either wet or dirty, it is still

necessary to use aflush gasorpurge gasto keep liquids and solids away from the

inner seal. The flush gas is supplied between the inner seal and a seal housing

labyrinth seal. It is important to consider that a reduction of nitrogen pressure below

the sealing pressure will result in process gas emission and possible damage to the

inner seal faces, so some back-up or safety provisions may be needed to avoid these

consequences (see Barrier gas and flush gas supply systems for double sealsand

Shutdown Protection Considerations). If the nitrogen supply is known to have poor

reliability, a tandem seal arrangement may be preferable, especially if protection

strategies are burdensome. Since nitrogen is not always available at high enough

pressures, double seal arrangements are usually limited to lower pressure services

such as FCC or coker wet gas.

The tandem seal is the most commonly used arrangement on compressors,

especially in moderate to high pressure services. The dual seals are oriented in

tandem to limit outward leakage (see Figure 200-52), with the inboard (primary) seal

normally seeing essentially all of the sealing differential pressure. In this

arrangement, seal gas is supplied between the inboard seal and a seal housing

labyrinth, as shown in Figure 200-53. In addition to acting as the back-up seal to the

primary seal, the outboard (secondary) seal contains all but a fraction of the primary

seal leakage under normal conditions. The cavity between the two seals is typically

vented to flare (or safe location) throughprimary ventporting in both the seal

housing and compressor. If the seal gas is environmentally harmful, or the process

gas is toxic, a tandem seal with an intermediate(or interstage) labyrinthshould be

employed, along with a nitrogen buffer gas. The intermediate labyrinth is located in

the interstage cavity between the primary and secondary seals, so pressure in this

cavity is normally low enough to easily allow plant nitrogen to be used as buffer

gas.

Buffer gas enters the interstage cavity through a port between the labyrinth and the

secondary seal and flows across the labyrinth, preventing seal gas from reaching and

leaking across the secondary seal. Most of the buffer gas exits the seal through the

primary vent where it mixes with leakage from the primary seal, while a smaller

amount leaks across the secondary seal. The tandem seal arrangement generally

requires the most extensive auxiliary system, which must deliver seal gas, deliver

buffer gas (if needed), and monitor seal venting conditions. Since the tandem

arrangement allows for a larger variety of gases to be considered as seal gas(including process gas from the compressor discharge), it is currently the only

arrangement choice for moderate to high pressure services. An external seal gas

supply is sometimes needed when process gas can not be appropriately treated, or

when compressor discharge gas is insufficient due to low rotational speed of the

compressor (starts, stops, idle time).

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A seal housing labyrinth seal, just inboard of the dry gas seal assembly is often

included in the design of any of the above three arrangements (see Figure 200-53).

This labyrinth:

limits the amount of seal gas or flush gas flowing into the compressor

provides a high velocity area for flowing seal gas or flush gas to minimize the

chance of solids and liquids from getting close to the dry gas seal

limits leakage to atmosphere in the event a primary seal failure

Fig. 200-52 Simplified Tandem Seal Arrangement with Intermediate Buffered Labyrinth

(Courtesy of Flowserve Corporation)

PROCESS ATMOSPHERE

Inert Buffer GasLeakage

Inert Buffer

Gas

Intermediate labyrinth

Fig. 200-53 Simplified Tandem Arrangement Showing Shrouded Seal Face Design, Primary Seal Labyrinth,

and Separation Gas Arrangement (Courtesy of Flowserve Corporation)

Seal GasPrimary Seal Vent

Inert Separation GasSecondary

Seal Vent

Primary Seal

LabyrinthSeparation Gas Labyrinth Seal

Seal Face Shrouding

Seal HousingLabyrinth

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The seal housing labyrinth seal can either be integral to the seal assembly or

provided as a separate compressor component. Similarly, labyrinth seals can be used

on the outboard side of the seal assembly to prevent bearing lube oil from

contaminating the seal faces. (This and other options are described in better detail in

Separation Seal.) For either application, the use of abradable seals(rotating

labyrinth teeth running within a soft, non-metallic, close-clearance stationary ringinsert) should be avoided, as some users have experienced failures due to excessive

heat generation and particulates originating from the abradable material. More

recent abradable seal offerings using a thin layer of abradable coating on metal

inserts appear less likely to cause such problems, but these designs should be

carefully reviewed, and the review should include a check of installed experience.

Properly engineered abradable seals continue to be acceptable for interstage and

balance piston sealing duty.

Seal Faces

Seal face materials and designs vary between different suppliers. Since the seal

faces are the components that have the greatest influence on the reliability and

leakage performance of the seal, they are the focus of ongoing designimprovements. Face designs must be optimized to address numerous issues,

including:

Hydrostatic lift (slight separation of the faces caused by pressure while rotor is

static)

Low speed contact tolerance

Dynamic lift-off properties

Gas film stiffness variations

A range of seal gas properties and their variability

Stresses and deflections due to sealing conditions (pressure and temperature)

Stresses and deflections due mounting/driving forces and dynamic forces Tolerance to reverse rotation

Some of these issues are addressed withface materialselection, which is often

dependent on the manufacturer, but is sometimes driven by service conditions. Most

of the installed population of dry gas seals use tungsten carbide for the rotating seal

face and carbon for the stationary seal face. Silicon-based materials have been

gaining favor over the years for both rotating and stationary faces, and especially for

the latter in high pressure services where carbon materials deflect excessively. At

present, and depending on the supplier, rotating faces are typically tungsten carbide

or silicon nitride, while stationary faces are either carbon or silicon carbide. Since

silicon carbide can not provide the very good low speed touch tolerance of carbon, it

must be coated with diamond-like carbon (DLC), a very hard and low frictioncoating that is also used on computer hard disks, in order to protect the seal faces

from low speed touch damage. Other material combinations have been used,

especially with unusual or extreme conditions. As with many machinery

components, it is important to recognize the potential for numerous variations of

generic material families, as they can be significant with regard to properly meeting

service condition requirements.

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Rotating face materials are generally very brittle, and are thus prone to break up

with little warning in adverse conditions. Since the rotating face materials are also

very hard, loose, broken fragments can cause damage to the seal. In some failures

with tungsten carbide faces, which tend to break up into relatively large fragments,

damage has extended to the compressor. In order to mitigate damage or unsafe

conditions in the event of a failure, a shroudedface design (see Figure 200-53) isrequired, even for silicon carbide-based materials, which can still cause damage to

the seal despite their reputation for breaking up into harmless powder. The brittle

properties of the rotating seal faces also necessitate compliant centering devices for

the faces as well as the seal assembly shaft sleeve. This protects the faces from

excessive stresses caused by radial differential thermal growth between the shaft,

seal sleeves and faces. O-rings are sometimes used for these compliant centering

devices, but in many cases, metallic devices are needed.

Seal face groove geometryalso varies between suppliers, and has evolved over the

years. Most suppliers offer both unidirectionaland bi-directionalface designs but

have tended to specialize in one or the other. Unidirectionalfaces typically have a

spiral groove geometry, although L-shaped grooves have also been used (see

Figure 200-54). Unidirectional seals directionally offer better performance with

regard to lift-off, gas film stiffness and stability, but this differential in performance

appears to have seldom precluded the successful application of a bi-directional

design in most applications. Unidirectional seals have the disadvantage of being

intolerant of reverse rotation, which can cause contact damage to the faces. Because

of this vulnerability, it is important to incorporate assembly features (both labeling

and geometry differences) which can help prevent the installation of the wrong seal

parts or assemblies (inboard vs. outboard) on a between bearings (double ended)

compressor design.

Fig. 200-54 Unidirectional Seal Face Groove Geometry Courtesy of Flowserve Corporation

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Bi-directionalface designs have a greater variety of groove geometries among

suppliers, including U-shapes, spruce tree-shapes and T-shapes, (see

Figure 200-55). with their symmetry offering equal performance in both directions

of rotation. In some low gas density applications, gas film properties in a bi-

directional design may be prove marginal or insufficient, in which case a

unidirectional design should be considered. Note that such an assessment is mostmeaningful when comparative numbers and criteria are provided, and especially

when it comes from a supplier that normally provides a bi-directional design. Due to

the ability to handle reverse rotation, bi-directional designs are somewhat more

desirable in services where compressor flow reversal potential exists (i.e., back

pressure services rather than recycle services), especially if there is a history of

compressor discharge check valve problems. Bi-directional designs also require

only one seal assembly design for both sides of a double ended machine, alleviating

the need for labeling and mis-installation provisions. The dual compatibility also

offers additional maintenance flexibility with regard to spare seal change outs and

repairs, but it should not result in the reduction of the seal sparing to anything less

than 100 percent (see Maintenance Considerations).

Secondary Sealing Elements

Secondary sealing elements (different than the secondary seal in a tandem

arrangement) provide sealing between the dry gas seal assembly (or cartridge) and

the compressor, as well as between various seal components. Typically, elastomeric

o-rings are used as the secondary sealing elements, although other seal types are

used to address specific problems. Most secondary sealing elements are static (once

parts are assembled, there is no movement of the parts that form the joint). As with

other machinery applications, it is important to select materials that are compatiblewith the normal and potential gas streams seen by the seals. In addition, high

pressure applications must be evaluated for the potential of extrusion and explosive

decompression(the latter is a function of sealing pressure, gas composition and

compressor system decompression rate). High pressures may require the use of high

Durometer elastomers or polymer (such as PTFE) materials. The polymer seals

typically use metallic springs to provide the proper contacting or energizingforce.

Fig. 200-55 Bi-direction Seal Face Groove Geometry (Courtesy of Flowserve Corporation)

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Besides a higher pressure capability, polymer seals also offer much longer (if not

infinite) shelf life than elastomeric seals. Given the requirements to replace

elastomeric seals on stored seals after several years of storage, gas seal designs that

use polymer seals may be worth considering in order to avoid this additional

maintenance which typically must be done at the seal manufacturers facility.

Installation and maintenance should always be considered in the secondary sealing

element joint design, especially those between the seal housing and compressor

casing, and the seal sleeve and shaft. Optimum o-ring placement and tapered

diameter changes can minimize or eliminate the potentially damaging action of

sliding o-rings across components during installation, as well as reduce potential for

o-rings falling out of ID grooves during installation. Special tools for the installation

and removal of dry gas seals are also an essential part of protecting these critical

components. On new compressor installations, these special tools should always be

pre-tested and used in the compressor suppliers factory, as they have frequently

been designed improperly or ignored, resulting in damage to the seals during

installation and inspection activities.

In addition to static secondary sealing elements, there are also dynamicsecondarysealing elements, which seal the moving joints between the stationary seal faces and

their retainers or housings. The dynamic secondary sealing element must allow the

stationary seal face to move axially in order to accommodate lift-off, gas film

thickness changes and both axial movement and thermal growth of the rotor, while

at the same time providing gas containment and resisting extrusion. These

conflicting requirements make dynamic secondary sealing elements the second most

critical components in a dry gas seal (after the faces). As such, the secondary

sealing elements should always receive appropriate attention when selecting the seal

design as well as inspecting the seal, both following shop testing and field

operation.

Wear out and axial sticking (hang-up) of the dynamic secondary seal can result inexcessive leakage, and hang-up can also result in damaging face contact. Potential

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Cmp en 200cvx