This document compares screw compressor and centrifugal compressor options for a chlorine gas compressor. Centrifugal compressors generally have higher efficiencies but lower turndown capability compared to screw compressors. For this application, a centrifugal compressor could provide annual electricity savings of $147,000 due to its lower power requirement. However, screw compressors are better suited if large swings in gas composition are expected, as centrifugal performance is more sensitive to changes in gas properties. Maintenance is easier for options with vertically-split casings. The document provides detailed technical considerations and specifications from multiple compressor vendors for each compressor type.

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SLIC HCL RECYCLE PROJECT
TPT PROCESS TECHNICAL NOTE
SCREW COMPRESSOR VS CENTRIFUGAL COMPRESSOR
FOR CHLORINE COMPRESSOR K6570
TN Number : 1975Z-065-TN-C312-0001
Prepared by : Guo Rong PENG (Process Engineer) & CangTo CHEAH (Senior Rotating
Equipment Engineer)
Reviewed by : Sern Hoe CHEAH (Process Lead) & Xin Cong WANG (Rotating Equipment Lead)
Date : 1st
issue – 15th
Jan 2015
1.0 INTRODUCTION
Based on Design Note GEN2-1-ROTE-RP-001 Rev A1 (SLIC-ROTE-03), EPC contractor
shall investigate the use of centrifugal compressor for Chlorine Compressor K6570,
compared to the screw compressor as specified by Licensor during FEED phase.
TPT Process and TPT Rotating have made inquiry to the centrifugal compressor and
screw compressor suppliers to gather the design information and pros/cons regarding
each type of compressor.
The following section lists the design information from the compressor suppliers. The
selection of the type of the compressor will be based on the comparison of the design
information as supplied by the compressor suppliers.
2.0 COMPARISON BETWEEN SCREW COMPRESSOR AND CENTRIFUGAL
COMPRESSOR FOR K6570
Vendor 1 (dry screw) Vendor 2 (centrifugal) Vendor 3 (dry screw) Vendor 4 (centrifugal)
Equipment in Skid 1. Compressors
2. Gearbox
4. Electric motor
3. Ladders & platforms
4. Process gas suction
strainer
5. Inter-stage cooler
6. Discharge cooler
7. Bypass valve
8. Noise & attenuation
9. Lubrication system
10. Base frame
11. Control panel
12. Seal oil system
1. Compressor
2. Gearbox
3. Electric motor
4. Dry gas seal panel
5. Lube oil piping
6. Base frame
7. Lube oil system
including rundown tank
8. Process gas inter-
stage cooler
9. Process gas after-
cooler
10. Control panel
11. LP and HP anti-
surge valves
1. Compressors
2. Gearbox
4. Electric motor
4. Process gas suction
strainer
5. Inter-stage cooler
6. Discharge cooler
7. Bypass valve
8. Suction & discharge
silencers
9. Lubrication system
10. Base frame
11. Control panel.
12. Discharge non-
return valves
13. Dry gas seal system
1. Compressor
2. Planetary gearbox
3. Electric motor
4. Dry gas seal panel
5. Lube oil piping
6. Base frame
7. Lube oil system (1
shaft driven pump, 1
electric motor driven
pump)
8. 2 x inter-stage
coolers
9. 1 x bypass cooler
10. Control panel
11. Anti-surge valves
Compressor Stage 2 stages in 2 compressor
casings.
2 stages within one
compressor casing
(barrel type casing).
3 stages in 3
compressor casings.
3 stages in 1
compressor casing.
Efficiency of Each
Stage
1
st
stage polytropic
efficiency is 73.64%, 2
nd
stage polytropic efficiency is
63.31%.
1
st
stage polytropic
efficiency is 83.6%, 2
nd
stage polytropic
efficiency is 82.5%.
Not available Not available
Compressor casing
construction
Vertically-split Vertically-split Vertically-split Horizontally-split
Compression
Power of Each
Stage / kW
1
st
stage compressor
absorbed power is 607 kW
at normal condition, 2
nd
stage compressor absorbed
power is 737.9 kW at normal
condition. Total
compression power is 1345
Shaft power is 849 kW
at normal condition.
Total compression
power 1176 kW.
Total compression
power 945.1 kW.
Note: Compression
power can be reduced
(i.e. increase of
polytropic efficiency) by

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Compressor performance:
The centrifugal type shown better performance in term of efficiency but it is limited to turn
down capability (as lowest threshold of suction volume flow rate in centrifugal
compression is restricted by compressor surge limit). The volumetric efficiency of dry
screw compressor is a function of rotor slip; this is the internal leakage from the higher
pressure to the lower pressure end (high pressure gas leaks through the gap between
meshes of male and female rotors); hence reducing volume capacity of the compressor.
If large turndown is not envisioned in the plant, it is advisable to consider centrifugal
compression from OPEX and environmental perspectives (i.e. less demand on electricity
consumed by compression process, which indirectly reduces the contribution of
greenhouse gases to the environment). The power saving of centrifugal compression
system is approximately 327 kW (difference between vendor 3 and vendor 2); this yields
an annum saving of 2452500 kWh (based on 7500 hours on-stream factor). When
multiplied with nominal electricity rate (0.60 Chinese Yuan/kWh), this is equivalent to 1.47
million Chinese Yuan of electricity saving per year.
kW. optimizing the flow
coefficient. Current flow
coefficient of 1
st
impeller is 0.0457,
which is not optimum
according to Ludtke’s
compressor literature
[4], optimum flow
coefficient to attain
good efficiency is
approximately 0.08.
Impeller diameter to be
reduced in order to
achieve higher
efficiency; additional
impeller is needed to
compensate polytropic
head loss due to
reduced impeller
diameter. To be
discussed with supplier
at DED stage, if
needed.
Total Power / kW Motor rated power is 1500
kW.
Not available Motor rated power 1350
kW.
Motor rated power 1050
kW
Cooling Water Duty
/ kW
53.6 m
3
/hr for inter-stage
cooler and 20.8 m
3
/hr for
lube oil cooler
6030 kg/hr for lube oil
cooler
Not available 12 m
3
/hr for lube oil
cooler
Chilled Water Duty
/ kW
105.4 m
3
/hr for inter-stage
cooler
Not available Not available Not available
Skid Dimension 9m(W)×11m(L)×5m(H) 5.72m(W)×9m(L)×5.46
m(H)
Not available 3.7m(w)x9.8m(L)x5.3m(
H): for compressor and
gearbox only.
Weight 57300 kg. 75200 kg Not available 16500 kg (compressor
and gearbox only)
Maintenance
(Special Tools)
Not available Not available Not available Not available
Manufacturing Time 60 weeks from receipt of
Instruction to proceed
Not available Not available Not available
Reference list 3 reference projects in
Chlorine gas application, but
none of them have similar
duties in term of pressure
and compression power.
27 reference projects in
Chlorine gas
compression. 7 of them
have similar duties in
term of pressure,
compression power,
and number of
compressor casing.
12 reference projects in
Chlorine gas
compression. 1 of them
has similar duty in term
of pressure and
compression power;
however it only consists
of two casings and run
at much lower rotational
speed compared to the
proposed solution for
SLIC project.
6 reference projects in
similar Chlorine gas
compression in term of
pressure, compression
power and number of
compressor casing.

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Requirement of high precision timing gears:
Note that rotations of male and female rotors of dry screw compressor are transmitted by
a set of high precision timing gears housed in the bearing compartment; the timing gears
have to be of good quality in order to maintain the timing of rotors (avoid rotor clashes)
and minimize machinery noise. Operating the screw compressor at off-design condition
(e.g. lower or higher compression ratio) would cause gas fluctuate / pulsation between
process discharge piping and compressor discharge nozzle (Pulsation levels are
minimized when the compressors operate near the design conditions, but can be
significantly increased at off-design conditions.). Compressor will rotate in reverse
direction on gas backflow, therefore minimizing gear backlash is important (this can be
controlled by split-driven gear where the female rotor’s timing gear is made to be movable
relative to its hub; which requires careful adjustment during start-up / maintenance).
If the clearance between rotors is larger than design value, the volumetric efficiency will
be reduced as high pressure gas leaks back to low pressure end through the clearance. If
the clearance between rotors is smaller than design value, it may cause rubbing of rotors
and would reduce the useful life of dry screw compressor (leading to unplanned shutdown
is credible scenario).
The centerline between male and female rotors shall therefore be maintained within tight
tolerance in order to avoid potential clashes between the rotor meshes; it means careful
design and control of lube oil system are needed to ensure precise hydrodynamic lifts are
created between journal bearings and rotating shafts at both loaded and unloaded
conditions.
Compressor train arrangement:
The dry screw compression of vendor 1 requires two separate casings. Two casings
solution coupled with single driver can either be:-
a) Driven by an electric motor of single output shaft with a gearbox of two output shafts.

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b) Or driven by an electric motor of single output shaft with gearbox of one output shaft,
and inter-connecting coupling between two compressor casings.
c) Or driven by an electric motor of two output shafts and two gearboxes.
For vendor 3 (dry screw), it requires 3 compressor casings with the following train
arrangement (motor of 1 output shaft with gearbox of 1 bull gear and 2 pinions):

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Any of the above cases (for dry screw compression) require careful design in term of rotor
dynamics, multiple shafts length (motor shaft, gearbox shafts, couplings, 2 x compressor
shafts of low pressure casing, 2 x compressor shafts of high pressure casing) may incur
significant challenges (during detailed engineering phase) on turbo-machinery lateral
critical speeds as well as torsional vibrations and stresses.
The centrifugal compression can be executed within one compressor casing (where
impellers are arranged in back-to-back configuration to minimize unbalanced axial thrust
due to compressed gas as well as reducing the required pressure for primary seal gas;
the latter is achieved by arranging the suction volute of each compression stage at the
bearing ends). The centrifugal compression of single casing solution will be coupled to
the main driver via a gearbox; where the shaft length (motor shaft, gearbox shafts and 1 x
compressor shaft) is minimized (as compared to two and three casings solution). This is
preferable from turbo-machinery rotor dynamic perspective.
Adaptability of compressor with respect to change of gas compositions:
The performance curve (polytropic head vs. suction volume flow rate) of centrifugal
compression is sensitive to the gas compositions of medium subject to compression. The
change of gas compositions will alter the slope of abovementioned curve:
Figure extracted from “Compressor Performance Aerodynamics for the User” [2]

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If significant change of gas compositions is expected in the process stream (assume
volume flow rate enters compressor suction nozzle is maintained constant); the slope of
polytropic head (H) vs. suction volume flow rate (Qdot) of centrifugal compressor will
change accordingly. The change of slope (denoted by blue color curve) will alter the
impeller polytropic head, hence affecting the discharge pressure as well as discharge
volume flow rate of first impeller.
As the discharge volume flow rate of first impeller changes; operating point of subsequent
impeller will not only be affected by the new slope (due to change of gas composition) of
second impeller, it will also be affected by off-design inlet volume flow rate (delivered by
the upstream impeller); this is so-called aerodynamics mismatch [3] where both polytropic
head and polytropic efficiency are affected by the off-design suction volume flow rate.
If significant change of gas compositions is expected during normal compression duty,
the overall compressor performances (i.e. polytropic head, polytropic efficiency,
compression power, overall surge limit, and overall choke limit) will be affected.
Whereas positive displacement (dry screw, wet screw, reciprocating, diaphragm, etc.)
compressors can tolerate huge swing of gas compositions provided actual suction volume
entering the compressor is maintained within the design limits. It is advisable to consider
screw compressor if significant changes of process gas compositions during normal
operation is envisaged.
Nitrogen gas start-up case:
Nitrogen gas start-up duty for the chlorine compressor had been specified (suction
volume flow rate to be defined by compressor supplier) and forwarded to compressor
suppliers to ensure it is feasible to start-up with nitrogen gas. Both dry screw and
centrifugal compressions are able to start-up with nitrogen gas, as tabulated below:-
Vendor 1 (dry screw) Vendor 2 (centrifugal) Vendor 3 (dry screw) Vendor 4 (centrifugal)
Suction pressure 2.913 barA 2.863 barA
(downstream of suction
throttling valve)
2.913 barA 2.913 barA
Discharge pressure 12.7 barA 6.4 barA (Note 1) 12.7 barA 6.4 barA (Note 1)
Required mass flow 14874 kg/hr 10800 kg/hr 15106 kg/hr 12491 kg/hr
Compression power 1162 kW 340 kW 983 kW 368 kW
Discharge
temperature
100 deg. C 51 deg. C 95 deg. C 40 deg. C
Note 1: The centrifugal compressor cannot achieve 12.7 barA discharge pressure (as
required in FEED data sheet) due to the fact that it is driven by fixed speed electric motor.

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If discharge pressure higher than 6.4 barA is needed, the suction pressure should be
increased accordingly (suction pressure of approximately 5.6 barA will yield 12.7 barA
discharge pressure; to be investigated in more depth with compressor supplier and the
development of plant start-up philosophy at detailed engineering phase).
Ease of maintenance:
Screw compressors (vendor 1 and 3) and centrifugal compressor proposed by vendor 2
are constructed with vertically-split outer casing, which means the process piping
connections need not be dismantled when the compressor internal / rotor to be pulled out
from the casing during maintenance. This will also ensure the tightness of process
flanges with compressor nozzles be maintained as it is before the compressor is shut-
down for maintenance.
However, horizontally-split outer casing is offered by vendor 4; this will complicate the
maintenance activities as the process piping need to be dismantled prior to the removal of
compressor internal.
Acoustic pulsations [1]:
Dry screw compressor generates high frequency pulsations that move into the process
piping and can incur acoustic vibration problems (similar to the type of problems
encountered in reciprocating compression). Screw compressors generate pulsation at
multiples of the pocket-passing frequency, which is defined as the number of lobes on the
male rotor multiplied by the compressor running speed.
Pulsation can result in increased vibration and fatigue failures of small-bore piping and
instrumentation attached to the discharge silencer and to the piping downstream of the
silencer. The vibration levels can be further increased if the pulsation frequency is
coincident with one of the mechanical natural frequencies of the small bore piping or
instrumentation.
The acoustic pulsations in reciprocating compression can be reduced by means of
volume bottles; while the dry screw compression would require compressor
manufacturer-supplied proprietary silencers (pulsation frequency in dry screw
compression is higher than reciprocating compression).
Although silencers of dry screw compressor are designed to attenuate pulsation at certain
frequencies, they can also amplify pulsation when the excitation frequencies are
coincident with the acoustical natural frequencies of the silencer itself. Pulsation in the
silencers can also increase the vibration levels of the compressor rotors and can cause
electrical problems to be fed into the local bus.
Commingle of nitrogen gas into process stream:
Self-acting dry gas seal is recommended (as offered by vendor 2 and 3) for chlorine
compressor [5]; it is used where leakage of the process fluid is not allowed, and
consumption of filtered buffer gas (typically nitrogen gas) needs to be minimized.

8. Page | 8
Schematic of double opposed self-acting dry gas seal
Each screw compressor casing consists two rotors (namely male and female rotors),
which leads to the requirement of four dry gas seal cartridges per casing. In the case of
vendor 1 and 3, eight and twelve dry gas seal cartridges are required, respectively.
Centrifugal compressor consists of one shaft; two dry gas seal cartridges are needed for
each centrifugal casing. This implies that less sources of nitrogen gas (as buffering gas of
dry gas seals) will commingle with process gas if centrifugal solution is implemented.
Vendor 1 proposed seal oil system, where further treatment of seal oil (oil + chlorine gas)
is needed.
Vendor 4 proposed labyrinth seals, which is an old technology. Labyrinth seal is made up
of a number of evenly spaced thin strips or teeth, the diametral clearance between each
of the seal teeth and the rotating shaft is equivalent to a series of orifices; leakage rate is
proportional to this clearance [5].
Sectional view of labyrinth seals proposed by Vendor 4
Labyrinth seals have higher leakage rate (clearance of labyrinth seals with rotating shaft
increases when rubbing occurs due to high vibration, this is so-called mushrooming effect
as depicted below. High vibration amplitude “Ac1” is expected when the compressor
passes through first lateral critical speed as shown in rotor response plot below [6], at
start-up or coast down) compared with self-acting dry gas seals; the seal faces of the
latter technology are mechanically-controlled by springs which are evenly distributed
around the circumference of stationary rings, i.e. leakages through self-acting dry gas
seal are therefore minimized.

9. Page | 9
New and clean labyrinth seal: Turbulence creates resistance to leakage flow [5]
Rubbed labyrinth seal (i.e. mushrooming effect): Clearances increased, turbulence
decreased, and leakage increased [5]

10. Page | 10
Rotor response plot (extracted from API 617, 7th
edition) [6]
Mechanical wearing parts:
Mechanical wearing parts in dry screw compressor are: timing gears (two per casing),
thrust bearings (two per casing) and journal bearings (four per casing). In centrifugal
compressor, mechanical wearing parts are identified as follows: thrust bearing (one per
casing) and journal bearings (two per casing).
The increased number of mechanical wearing parts in dry screw compressor is likely to
reduce the availability and “mean time between failures” of the compression system.
Refer to “Availability and MTBF” section for more details.

11. Page | 11
Availability and MTBF:
Equipment availability and mean time between failures based on 8760 hours/year
(Jan ’99 Hydrocarbon Processing Magazine) quoted by vendor 1:
Equipment Availability Mean time between failures (years)
Best Average Best Average
Centrifugal 0.999 0.997 15 8
Dry screw 0.996 0.990 8.0 4.0
The above figures are derived per machine basis; where the availability may be further
reduced if multiple machines are configured in series (i.e. dry screw compression).
Overall scoring table:
Where: 0 low rank / undesired, 1 high rank / desired.
Parameters Dry screw Centrifugal Remarks
Compressor
performance
0 1
Centrifugal compressor saves more
energy.
Requirement of high
precision timing gears
0 1
Backlash of timing gears due to
backflow needs to be avoided in
dry screw compressor.
Compressor train
arrangement
0 1
Lengthy shaft arrangement of 2
and 3 casings solution for dry
screw compression not preferable
from rotor-dynamic perspective.
Adaptability to change
of gas compositions
1 0
Screw compressor has higher
flexibility to accommodate change
of gas compositions.
Nitrogen gas start-up
case
1 0
Both screw and centrifugal
compressors can handle the
specific start-up duty with nitrogen
gas. However, suction pressure for
centrifugal solution needs to be
increased if higher discharge
pressure is required (to be checked
with plant start-up philosophy at
detailed engineering).
Ease of maintenance 1 1
Screw compressors (vendor 1 & 3)
and centrifugal compressor from
vendor 2 are built with vertically-
split outer casings; dismantling of
process piping is not required
during compressor maintenance.
Note: Vendor 4 offers horizontally-
split outer casing; not preferred
from maintenance perspective.
Acoustic pulsations 0 1
Acoustic pulsations will create
excitation frequency on the directly
attached process piping (flange
connections prone to gas leak) and
structure borne energy that can
excite components of nearby
systems. Screw compressors
prone to acoustic pulsations due to
its nature of compression
mechanism, i.e. positive
displacement.

12. Page | 12
Parameters Dry screw Centrifugal Remarks
Commingle of nitrogen
gas into process
stream
0 1
Screw compressor requires more
dry gas seal cartridges per casing;
therefore the amount of nitrogen
gas dilution with process stream is
higher than centrifugal solution.
Mechanical wearing
parts
0 1
More mechanical wearing parts
identified in dry screw compressor
(more spare parts and
maintenance activities); contribute
to its low availability and MTBF
indexes.
Availability and MTBF 0 1
Centrifugal compression has higher
availability and MTBF index
compared to dry screw
compression.
Reference in similar
duties
0 1
The centrifugal compression has
many references with similar
pressures, number of casing, and
compression power.
The 3 compressor casings solution
for dry screw compression
(proposed by vendor 3) has no
similar reference in term of
pressure and compression power
range.
Overall score 3 out of 11 9 out of 11
Based on the above scoring table, centrifugal compression is recommended from OPEX,
technical, environmental, similar running-unit experiences and operating safety
perspectives. Please refer to Procurement’s commercial note with regard to CAPEX
comparison of dry screw and centrifugal compressions.
References:
[1] Pulsation, vibration, and noise issues with wet and dry screw compressors;
Proceedings of the fortieth turbo-machinery symposium (September 12-15, 2011 Houston
Texas).
[2] Compressor Performance Aerodynamics for the User 2nd
Edition; by Theodore Gresh
(March 2001 Elsevier Science & Technology Books)
[3] Hydraulic shop performance testing of centrifugal compressor; Proceedings of the
thirty-fourth turbo-machinery symposium, by Gary M. Colby (2005, Houston Texas)
[4] Process centrifugal compressor: Basics, function, operation, design, application 1st
Edition; by H. K. Ludtke (2004, Springer)
[5] Compressor seal selection and justification, Proceedings of the thirty-second turbo-
machinery symposium, by Stephen L. Ross and Raymond F. Beckinger (2003, Houston
Texas)
[6] API 617 Axial and Centrifugal Compressors and Expander-compressors for Petroleum,
Chemical and Gas Industry Services, 7th
Edition (Reaffirmed January 2009, American
Petroleum Institute)