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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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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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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.)
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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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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 multi- stage 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---------------------------------------------------------------------=
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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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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