1.

Auxiliary systems - Proven solutions

 Lube Oil System
 Sealing system
 Thrust balancing
 Capacity control
 Surge control
2. Maintenance Plan- Trouble shooting
3. Revamps and retrofit
4. Operation of centrifugal compressors
5. Performance Evaluation
Auxiliary systems -Introduction
 For safe and reliable operation of compressor, following
auxiliary systems provide support for compressor
operation.
 Lube oil system: To take care of the lubricating
requirements of the compressor, driver and gear box
 Seal oil system : to provide effective sealing of
compressor by either floating ring seal system or
mechanical sealing systems
 DGS System: To provide compressor sealing by
employing process gas or a clean buffer gas.
 Axial thrust & Thrust bearings: To effectively
compensate the axial thrust and providing balancing of
gas thrust
Other Auxiliaries
Other auxiliaries also effect the health of compressors.
Some of the important are :
 Anti surge control system
 Control & Instrumentation
 Heat exchangers and cooling water system.
 Piping & Valves
 Electric motors, Gear boxes and air filters
Lube oil system
A pressurized oil system or systems shall be furnished to
supply oil at a suitable pressure or pressures as applicable,
to the following:
 The bearings of the driver, compressors and gearbox
 Continuously lubricated couplings.
 Governing and control-oil system.
 Shaft seal-oil system.
 Purchaser‟s control system wherever specified.
Design Basis
 Compressor lube oil systems are designed as per API 614,
Design activities involved are
 Selection of system e.g: combined LO-SO system, separate
seal oil system, Lube requirements for driver (ST or Motor)
and Gearbox. etc
 selection & sizing of various components of Lube oil
system.
 Making a P&I diagram, incorporating clients optional and
mandatory requirements of API , manufacture and shop
tests as per API
 Lube oil Pumps
 Lube oil pumps used are either centrifugal or rotary piston type.
selection criteria is generally dictated by customer requirements.
 Each pump is capable of supplying the normal oil flow required by
the equipment (Compressor, driver, gear box, gear couplings plus
greater of 20 percent of the normal oil flow or 40 l/min. Pump
sizing shall take care of Transient oil requirements.
 Pumps are sized to supply oil for increased bearing and seal
clearances. Accumulators may be provided to meet transient
control oil requirements wherever specified.
 An emergency oil pump is used to take care of coasting down
requirements of ST driven units.
 As a General practice, Turbine driven pump is considered for MOP,
and motor driven pump for AOP. For compressor units, shaft
driven MOPs are not acceptable and the pumps are externally
mounted.
 Lube oil Pumps
 A positive displacement pump gives more or less the
same oil flow at varying pressures depending on the
system's resistance. A positive displacement pump needs
a safety valve to take care of over pressure
 A centrifugal pump, however, has a delivery pressure,
which varies little on changing the flow established by
the system's resistance.

Rotory Pump characteristic Centrifugal Pump Characteristic

Base plates
Base plates are structural assemblies either made integral
with Lube oil tank or separate. For Gas turbine driven
units, the Lube oil tanks are built in the GT base plates.
In the case of ST driven units the Lube oil systems are
built over tanks and are single lift console type.

Lube oil Tanks

Tank sizing shall be done to fix working capacity ,
retention capacity and rundown capacities as defined in
API and customers specification
Lube oil OH Tank:
When specified, a separately mounted emergency lube-oil
rundown tank of SS material shall provide oil for the
coast-down period specified by the purchaser, which shall
not be less than 3 minutes of normal lube-oil flow.

Oil Accumulators:
An accumulator shall be provided to maintain the turbine
control-oil pressure during transients or to maintain lube-
or seal-oil pressure while the stand-by pump accelerates
from an idle condition to operating speed. The control oil
pressure shall be maintained above the equipment
manufacturer‟s minimum specified supply pressure at all
operating conditions.
Twin Oil Filters

 Twin full-flow filters with replaceable elements or

paper cartridges shall be provided. Filters shall provide
a minimum particle removal efficiency (PRE) of 90
percent for 10 micron particles and a minimum PRE of
99.5 percent for 15 micron particles, both as per ISO
4572
 The filters shall be located downstream of the coolers.
Compressor sealing systems
 The sealing of compressor shaft ends is accomplished by
either of the following methods.
 Labyrinth seals
• Oil seals
• Oil mechanical seals
• Floating oil seals
 Dry gas seals

 Selection of a particular sealing system for a compressor

depends basically on the gas medium and leakage quantity,
cost of the gas and nature of the gas like flammability and
environmental hazard etc.
15
Seal oil System

Sealing system for in line compressors

Floating oil seals Low pressure
Floating oil seals High Pressure
Combined LO/SO System Separate Sealing system
Seal oil Drain system
Dry gas seals
Areas of application
 Dry gas seals in centrifugal compressors find extensive
application in refineries, petrochemicals, fertilizers and
booster compressors. These are used to seal the
compressor ends from leaking the highly inflammable,
toxic and polluting gases to atmosphere.
 In refineries and petrochemicals, DGS are used for WGC
of FCC Units, and also in high pressure H2 recycle
Compressors, which are a vital part of HCU and DHDS/
DHDT units.
 In fertilizer industries, DGS are employed on CO2
machines and Synthesis gas and natural gas services
Selection criterion

Parameters to be considered for selection of DGS

 Sealing pressure / Settling out pressure
 Max temperature of the gas
 Size of the shaft
 Gas composition
 Allowable leakages
 Direction of rotation of the machine
DGS Working principle
A dry gas seal consists of
a spring-loaded
stationary ring (mounted
in the casing) held in a
retainer against a
rotating ring (mounted
on the rotor). A series of
profiled grooves are
made on the rotating
ring.

Gas, at a slightly higher pressure than the compressor ends

(usually process gas from the machine discharge or any
compatible gas available at higher pressure), called the
clean gas is introduced at the rings zone.
Working Principle contd…
 Due to the pumping action (aero-dynamic lift-off) of the
grooved rotating ring, a thin gas film of 3μ is formed
between the rotating and stationary rings. The ‘primary
leakage’ across the inboard seal is usually flared.
 A stable film of gas between seal faces keeps them
separate and enables them to run without touching each
other.
 Complete separation of sliding faces during operation is
absolutely essential because, contrary to liquid lubricated
mechanical seals, a gas lubricated seal would be unable to
dissipate the heat of friction.
 The seal faces have to be moved at a relative speed of 0.5
m/s to be separated at zero pressurisation. This is called
aero-dynamic lift-off. The gas film between the two sliding
faces acts like a spring.
Groove configurations
 Grooves on the rotating face, generate an aerodynamic
lift, which separates the faces and prevents wear.
 The groove profiles can be classified into two types
Unidirectional rotation
Bi-directional rotation
 Some of the typical groove profiles are shown below

Spiral grooves V grooves

Typical unidirectional groove profiles
 Dry Gas Seal Arrangements
Clean gas Primary Barrier
injection vent gas Clean gas Primary Secondary Barrier
injection vent vent gas

Single seal arrangement Tandem type arrangement

 Dry Gas Seal Arrangements
Clean gas Primary Secondary Secondary Barrier
injection vent injection vent gas

Tandem type seal with

Intermediate labyrinth

Clean gas Secondary Primary Barrier

injection injection vent gas

Double seal arrangement

Labyrinth Sealing systems

Typical CO2 Sealing System

Axial thrust – Selection of thrust bearing

 API 617 gives a guideline for the residual gas thrust

generated in a compressor.
 The resultant gas thrust to be limited to 10 % of thrust
bearing capacity.
 The total thrust coming on to the thrust bearing from
inboard and outboard couplings, gas thrust and
secondary thrust generated shall be limited to 50% of
the bearing capability.
 The resultant axial thrust has a direction and in back to
back compressor, it can be in either direction. A double
acting thrust bearing is recommended in a compressor.
 Axial thrust – how it is generated
 As a result of pressure difference across a stage Pd-Ps, an
axial force is generated towards the impeller inlet
 The cumulative thrust produced by all the impellers is
sizable and could adversely effect thrust bearing
 Thrust is generated in the following areas
 By Impeller stage
 By couplings
 From labyrinth seals
 The total thrust generated will be too large to be taken by
a practicable size of thrust bearing which can be seated in a
machine. To obviate this problem a compensating device is
placed on compressor rotor called „ Balancing Drum „
 Basics – Thrust

39
Thrust balancing of inline Compressors
Thrust generated by coupling
Thrust forces for a gear coupling are Caculated based
on the empirical formula given in API, whereas for
flexible-element couplings shall be calculated on the basis
of the maximum allowable deflection permitted by the
coupling manufacturer.

Where
Pr Rated Power KW
Nr. Rated speed
D dia under coupling mm
 Thrust balancing of inline Compressors
Gas thrust due to impellers ΣT = T1 +T2 +T3 + …
Balancing Drum thrust Tbal = Area of bal drum x DP
across drum

Coupling thrust F kg =

Where Pr Rated Power KW

Nr. Rated speed
D dia under coupling mm
Residual thrust = ΣT - Tbal
Total thrust = T res +/- F

Compressor Engineering 42
Thrust balancing of Back to Back Compressors
Thrust balancing of Back to Back Compressors
Gas thrust due to 1st phase impellers ΣTs = Ts1 +Ts2 +Ts3 + …
Gas thrust due to 2nd phase impellers ΣTd = Td1 +Td2 +Td3 + …
Residual Gas thrust ΣT = Σ Ts - Σ Td
Balancing Drum thrust Tbal = Area of bal drum x DP
across drum

Coupling thrust F kg =
Where
Pr Rated Power KW
Nr. Rated speed
D dia under coupling mm

Residual thrust = ΣT - Tbal

Total thrust = T res +/- F
Thrust generated in Labyrinth seal

80 bar 160 bar

I Phase II Phase

Detail of seal groove

Capacity Control

Capacity control can be achieved by the following

techniques:
1. Suction throttling
2. Variable IGV‟s
3. Speed control
4. Anti surge Recycling
5. Discharge throttling
Suction Throttling

 When single speed motor driven compressors have

varying operating situations like part loads, simplest
method of capacity control is by resorting to suction
throttling.
Suction Throttling
 A butterfly valve generally placed at the compressor
suction identifies either suction flow or discharge
pressure set point and puts the valve in throttling mode
Please see figure next sheet.
 The throttling is done such that the machine does not
enter surge zone. It is recommended to have a minimum
opening for suction throttle valve for motor driven
units.
 The second consideration is to ensure that the absolute
vacuum created at suction side near 1st impeller
should not pull lots of Lube oil from compressor ends
into the machines.
 The throttling optimizes the power consumed by the
driver for lower operating point.
Suction Throttling Curve 1

Y
Curve 1
Suction valve fully open, Curve 3
X design point
Curve 2
Curve 2
Suction flow m3/hr
Suction valve partially
closed Nx
Y lower operating point Ny

Curve 3 - Throttling line

Suction flow m3/hr

 Inlet guide vanes
 Adjustment of IGV‟s is a better compressor control compared to
suction throttling. According to the Euler turbo-machinery
equation the polytropic head is a function of the tangential gas
velocity at the impeller inlet (Cu1). By adjusting the inlet guide
vanes, the incidence angle at the impeller inlet can be altered
and thus the peripheral velocity changed.
 An important property of IGV‟s is to alter the impeller
performance. In principle, guide vanes can be provided at inlet
of each stage, but due to lack of space, multistage compressors
are normally equipped with IGV‟s at 1st compressor suction.
 Advantage of the adjustable IGV is its good part-load efficiency.
(lower than variable speed, higher than suction throttling) The
disadvantages being its Higher investment costs compared to
suction throttling
Speed control
 It is easier to effect flow control of variable speed
compressors driven by Stem or Gas Turbines by varying
the driver speed.
 This is done by taking a flow signal from the suction line
or discharge line and bring down the speed by transfer of
control to ST inlet control valve.
 This reduction can take place till the minimum governing
speed is reached. If the inlet flow further falls down the
machine will go into anti-surge bypass mode.
 The surge margin will move to left as the speed is
brought down.
Discharge throttling
 This control method restricts the pressure from the
compressor to match the required process pressure at
constant flow. Because the compressor is working harder
than the process requires.
 This control scheme is extremely inefficient, As a result
this control technique is rarely used.
Surge Principle

Case 1 rubber tank

Case2 Steel tank

Hand pump
 Surge in Centrifugal Compressors
 What is Surge ?
 Surge is a phenomenon of instability due to flow reversal in one
of the impellers or / compressors in a train.
 What causes a surge ?
 Discharge pressure or discharge resistance mounting high due
to Gas density increase ( steep increase in Mol. Weight )
 Compressor Inlet/ Discharge flow/ impeller Flow dropping
down due to blockages or due to drop in process consumption
 Excessive increase in m/c speed associated with reduced flow
rates or drastic drop in speed at high flow rates
 Classical solutions
 Anti surge Control
Provide Recycle bypass from discharge to suction or to vent.
 Performance Control
Reduction in Machine speed for a drop in throughput.
 Compressor surge and choke limits
Pressure

Surge point is that corresponds

to the lowest volume flow rate
Compressor Surge below which a compressor
Line
experiences instability
Anti Surge Line
Psurge
Design point is that which
represents an operating point
Design Point
for longer durations with
P1
optimum efficiency.

Choke point is that which

Quantity Bypassed corresponds to highest that can
pass through the stage, beyond
which the pressure developed
by the machine is very low,
fluid can not pass through the
channel
Q
req.
Q
surge
Q1 Flow
External causes and effects of Surge
• Restriction in suction or discharge of system
• Process changes in pressure, temperatures or gas
composition
• Internal fouling of flow passages of compressor
• Inadvertent loss of speed
• Instrument or control malfunction
• Malfunction of hardware such as variable inlet guide
vane, closing of suction strainers.
• Mal-distribution of load in parallel operation of
compressor
• Excess extraction of gases between sections
Basics Of Anti-surge Control system
Anti surge law h1 constant Variable speed
h1
K
P
Anti surge law h1 constant Motor driven
Constant speed Antisurge control
h1  K

Electric Motor driven

compressors
Anti surge law h1 / (P2-aP1) constant
h1
K
P2  aP1
Anti surge law h1 / P2 constant

h1
K
P2
Anti surge law h1 / (P2-P1) constant
h1
K
f P 
Anti surge law h2 constant

h2  K

Variable speed compressor with near

constant suction condition Not possible
to measure suction capacity
 About ASC design
 To design a reliable ASC system, We need
 Thermodynamic data for the compressor Suction, discharge
conditions at surge & guaranteed points, Compressor
performance maps & data sheets for different speeds with
predicted Surge points
 P&I diagram defining the process loop and required recycle
loops. instrumentation provided for the selected algorithm.
Details of instruments like range, type, etc. Anti-surge
controllers make, type and response time etc.
 Outputs:
 Definition of Anti surge control line, Selection of Anti surge
controllers and hardware. Review of P&ID, set points for the
controllers instrumentation and safety of compressor loop.
 Anti Surge Vale Sizing
 Draw start up, loading, Normal operations & shutdown
procedures.
Components effecting compressor health.

 Couplings
 Sealing area
 Wear out of seals
 Gaskets
 Rotor components like impellers
 Journal & Thrust Bearings
 Auxiliaries
Problem areas
A machine is considered (While it is running at rated
parameters) not healthy when one of the following
happen.
 Higher steam / power consumption
 Higher Internal recirculation
 External gas leakage
 Oil entry into compressor
 Blocked internals like diffusers impellers, seals
 High Vibrations
 Bearing failures
 Thrust bearing getting overloaded
 Frequent surging etc.
Important indicators for machine health
 Bearing Temperature Rise
 Drop in Capacity and head
 Drop in polytropic efficiency
 Fluctuation in speed, suction pressures etc.
 Rise in vibrations sound level and axial displacements
 Seal failure
 Seal oil ref gas delta P low
 HP seal oil drain too high
 LP seal oil flow excessive
 Oil entry into machine
 Gas leakage to environment
What factors affect compressor health ?
 Machines are long overdue for annual turn-around
 Machines are working with operational issues
 Visible higher fuel/steam consumption.
 Subsequent to an event like surging, very high vibrations
 Severe change in operating parameters/system resistance.

Failure can take place due

 Failure of rotating equipments
 Failure of auxiliaries
 Failure of control & Instrumentation
 Process upsets + one of the above
 Operator‟s responses
 Overload and Fatigue
 Rotating equipment failures - causes
 Design Defects.
 Non-conformance to design specifications and start up
procedures. Operational deviations and deficiencies.
 Manufacturing defects and Defects in installation.
 Incidents like foreign object entry.
 Poor preventive maintenance.
 Malfunction or failure of instrumentation and controls
and protective devices like vibration probe thermo-
element etc.
 Human Skill and lack of understanding of the machine
behaviour like ignoring alarms, wrong start up. Lack of
surveillance etc.
Goals of a good Maintenance Engineer
 Increased Availability.
By minimizing process upsets, anti surge system failures.
 Increased Reliability.
By addressing the most common failure modes and
bringing on-line data into the regular surveillance.
 Decreased Maintenance Cost & Lost Production Revenue.
By regular performance tracking and Proactive trouble
shooting to reduce the impact of the failures by
identifying root cause faster.
 Better Environment.
To reduce the risk to personnel and the environment in
highly toxic compression processes, such as, Nat gas re-
injection with H2S by having proper sealing system and
State of The Art.
Recommended maintenance Schedules
Sl
WHEN SIX BEYOND
N COMPONENT ANNUALY CHECK WHAT
STOPPED MONTHLY AN YEAR
o.

Journal Check wear and rub and also any sign

1 YES YES
bearings of oil starvation

Check wear and rub and also any sign

Thrust
2 YES YES of oil starvation
Bearings
Check axial movement of rotor
Check wear and rub and also any sign
of oil starvation
3 Oil seal rings YES YES
Check for healthiness of
O-rings
*Check after 3 years of operation for
erosion, corrosion , deposits and to
4 Rotor YES*
verify rotor unbalance, Replace if
mechanical damages, rubbing noticed
Check after 3 years for erosion,
deposits and corrosion ,replace worn
5 Diaphragms YES YES
out lab seals, Replace Gaskets and
clean return channels
Recommended maintenance Schedules
Sl WHEN SIX BEYOND
COMPONENT ANNUALY CHECK WHAT
No STOPPED MONTHLY AN YEAR

Check after 3 years for erosion,

6 Casings YES YES deposits and corrosion ,Replace
Gaskets

Check the alignment and compare

7 Alignment YES YES
OEM data

Replace cartridges if DP id higher

8. Filters YES YES than permissible, replace once in a
year irrespective of DP
Check oil levels and fill the levels if
there is a loss of level , clean the tank
9 Main oil tank YES YES
bottom if sediments are found and
when oil is replaced
*Check couplings,
10 Pumps YES* YES *check Cleanliness of inlet filters
Check pump internals once in 3 years
Recommended maintenance Schedules
Sl WHEN SIX BEYOND
ITEM ANNUALY CHECK WHAT
No. STOPPED MONTHLY AN YEAR

Check gear tooth contact, inspect

11 Gearboxes YES YES bearings Check for wear and inspect
oil spray pipe

Inspect hardware and membranes at

12 Couplings YES YES
every opening.

Check water side for deposits,

13 Coolers YES YES YES
Check transflow valves
Couplings
 Gear Couplings damages are found in equipments working in
process plants. Some of the main causes are:
 Loss of lubrication due to failure of LO system, due to sludge
formation in steam turbine driven units
 Shrink fitted couplings often slip due to low contact area or
low interference or damaged shaft tapers.
 Over load due to capacity up gradation or change in
operating parameters.
 Misalignment caused by piping forces or steam leakages.
 Produce higher axial thrust because of higher misalignment
and friction
 Dry flexible couplings provide solution to many of the
lubrication related problems and produce relatively less axial
thrust
Labyrinth seals
 High-pressure gas will somewhat leak back to the low-pressure
area through small clearances between the rotor and
stationary component This is called as “ leakage”. As this
lowers the efficiency of the compressor, a seal is used to help
stop the leakage flow and also to improve the long term
reliable operation.
 Every centrifugal compressor must have some means of
limiting or preventing gas leakage along the shaft where it
comes out through the casing. This is accomplished by using
various types of shaft seals.
 Labyrinth seal
 Restrictive carbon ring seals
 Non contact type floating ring seal, Mechanical contact seal
 Dry gas seal
Labyrinth seals
 Higher clearances increase leakages effecting
performance and lower clearances will increase
susceptibility of touching and groove formation.
 As a convention, labyrinth seals are made out of softer
material compared to shaft/impellers to reduce effect of
touching. Aluminum alloy is used for compressor seals.
Where Aluminum is not compatible SS strips are used
 As a solution more sophisticated seals are employed such
as turbine type, abradable and honeycomb type seals.
 The labyrinth seals are placed onimpeller eye, impeller
hub , Balancing drum, shaft ends and interstage
balancing drums
 O rings
 O rings failures do happen during assembly, due to sharp
edges and dents in mating parts
 Due to non compatibility with process gas in high
pressure applications such as explosive decompression
 They also fail due to wrong selection of O ring or groove
depth leaving very low or no gasket compression. Some
of the elastomers are suitable only for limited pressure
application and are likely to fail by extrusion for higher
pressures.
 Certain gases which are chemically active such as
Ammonia , special materials such as silicone elastomers
are employed.
 O rings
 An elastomeric high pressure seal is typically designed to
operate within ambient pressure to about 1,500 psi. At a
very high pressure, the seal must have sufficient strength
to resist extrusion into the clearance gap.
 Techniques to avoid extrusion in a high pressure seal
include decreasing the clearance gap, increasing the
elastomer modulus (Mod 100) .
 As a better solution for increased life of o rings in high
pressure applications, Teflon back up rings or complete
Teflon rings or spring loaded Teflon gaskets are used.
Bearings
 Increase of bearing clearances due to
 Excessive loading,
 Loss of lubrication,
 Dirty or contaminated oil
 Bearing selection not meeting
 Lateral analysis requirements
 Oil whirling
 Operational Issues
 Rotor dynamic excitation causing wear
 Surging
 Bearing Quality
 Improper bonding of white metal
 Wrong selection of bearings
Trends of parameters to be taken
Trends of critical parameters will help to recognize any critical
Parameter deterioration and initiate corrective actions .
-Reduction of load
-Reduction of speed
-Open recycle bypass
-Adjust utility conditions like cooling water ,N2, aux. steam etc

Inlet pressures Vibrations & Axial displacement

Inlet temperatures Lube oil pressures & Temperatures
Discharge temperatures Journal Bearing temperatures.
Discharge pressures Thrust Bearing temperatures.
Flow at inlet Inlet Seal oil pressures
Axial thrust Seal oil Delta P
Speed
Overhauls , Revamps & Retrofits of
Centrifugal Compressors
Turbo machinery overhaul
 Modern turbo machines can run reliably for many years, if
designed, applied & operated correctly.
 More important is to maintain the health of the machines
by techniques like on line monitoring & preventive
maintenance.
 With several years of continuous operation, it becomes
necessary to plan the periodical overhauls.
 Apart from planned stoppages, several unprecedented
emergency overhauls also are required to be executed.
 The basic reasons for which overhauls are taken up are
 Deterioration of performance of the machines.
 Excessive vibrations, axial displacements, To prevent a
major breakdown Rectify accidental damages
Objectives of Overhaul-Revamp

 To replace damaged /sick parts

 Investigate for performance related issues and
incorporate the findings/ recommendations of RCAs
 Clean up and clear fouling of gas path components
 Restore seal/bearing clearances for performance
improvement
 To incorporate state of the art like couplings and seals
 To synchronize upgrading / revamp
 Revamps & Retrofits - Introduction
 Machines with deteriorated performance due to
ageing/changed operating parameters limit plant‟s
growing needs.
 Solution to the problem is retrofit the existing
machines on techno-commercial considerations.
 OEM‟s capability and experience in the field is the
right source for customer to realise the solution.
 Short payback periods, operational debottlenecking
are primary considerations.
 Retrofits are different from overhauls while a retrofit
is planned during a turn-around.
 Considerations
 Capacity up-gradation of existing plants
 Performance deterioration over a period of time
 Incorporation of product developments/ state of art
technology
 Operational feed back and problems experienced
 Need for sustained reliability
 Replacement of obsolete products for want of regular
supply of spares and services
Benefits
 Reduction in Energy consumption & Capacity Up-gradation
 Meeting changes in operating parameters
 Service life extension
 Increase in reliability and ease of operation
Compressor Revamp Vs. Driver
 Compressor revamp often necessitates increase in
power, steam turbines can be modified to opening up
of additional nozzles.
 If head increase is required in compressor and
consequently speed increase, steam turbine poses a
limitation as increase in speed beyond Maximum
continuous speed is associated with increase in blade
stresses.
 Increase in power can be achieved by rewinding of
electric motor and increase in speed by changing
gearbox internals.
Drivers - Futuristic view
 When a driver is selected care must be exercised
 To forecast future revamp possibility and possible
increase in plant load.
 Deterioration in compressor inlet parameters.
 Providing higher power reserve in ST would make the
revamp more feasible and less expensive.
 When a driver of higher rating is selected initially,
considering future power requirement, present operation
becomes less optimum.
 If a driver of higher power rating has to be installed in
future, the space and area considerations would not allow.
 Revamp Requirements
Revamp required in
Purpose
Compressor Auxiliaries Auxiliaries Rotordynamics
Increase in Compressor Coupling & Valves & Check Rotor
through- internals Driver Antisurge Dynamics
put ( Gas Path ) control,
Cooler/Separator
Head Compressor Coupling & Antisurge control Check Rotor
increase internals Driver Dynamics

Dry gas Compressor Add DGS & Check Rotor

seals casing Sealing dynamics
compatibility system
Others CCC antisurge Dry flexible LEG type thrust Honeycomb
control for couplings bearings seals to
better control Greater Higher thrust increase rotor
reliability capability stability
Case study of a revamp
+ 10 to 20% Higher Capacity, + 10% higher head, +3% higher efficiency
 One or many of the above are customers requirements
 This results in the following :
 Higher number of impellers, higher size of impellers, different type of
impeller .
 Higher speed & Higher power demand from ST ( higher steam flow)
 Change the gear coupling size , Change gear ratio + Rewind the motor
 Critical speeds changes and encroached into operating zone.
 Increased impellers and size necessitates change of all the gas path
components and rotor becomes new. The factors to be considered are
 Time required to incorporate the changes,
 Downtime for incorporating changes,
 Cost of revamp components.
 Sum of all this is cost. Customer will decide for the revamp after
evaluating the ROI, Vi- a- Vis cost of acquiring new equipments
Options used for revamp
Startup/loading &Trouble shooting of
Centrifugal Compressor
 Compressor Operation

Precautions for startup/loading

 Every precaution is taken during startup and loading of
compressors. Some of the machines handling heavier
molecular weight gases like air, Propane, Propylene and
CO2, healthy starting procedures make the machine
operation smooth and trouble free.
 During design stage, designer examines the possible
mismatch between different stages and recommends
venting or recycling at appropriate stages which are
closed after the machine picks up rated conditions.
 Precautions for startup/loading
 Like wise, during trip, large quantity of gas is entrapped
in various sections . The high pressure zones create a
high resistance for the low pressure section forcing it to
go into surge.
 One should remove this pressure dam to enable the
machine smoothly coast down to rest. Anti-surge valves
and vent valves together remove this resistance.
 In case of seal failures one needs to evacuate the entire
trapped volumes for safety.
 The start up considerations are different for different
services. Some of these are detailed below
 Startup/loading :: Air compressors
 Motor driven compressors are always started with suction
throttling / IGV to ensure motor current does not exceed the
permitted value unless DOL start is envisaged for the unit
 Start the suction chillers after the machine reaches full speed.
Do not open cooling water fully to gas coolers till the pressure
build up in compressor exceed the cooling water pressure.
 The machine anti-surge/blow off valves are in “ open condition”
at start up. Wherever required open the bleed valves to reduce
the stage mismatch and choking ,
 For motor driven air compressors, provide a 2nd anti-surge valve
in parallel to prevent the machine to go into choking. Never run
the air compressors in closed loop
 Pressurize bearing chambers with external air, to prevent oil
being sucked into machine from ends, as the machine suction is
virtually brought to vacuum at start up,
Air compressor
Startup / Loading :: NH3 Refrigeration service
 During start up of NH3 compressors, do not open the cooling
water valves fully to prevent sub-cooling of vapors at inter &
after coolers. Open only after the machine reaches rated
conditions.
 Intermediate injections are not connected to respective stages
as the higher injection pressures force the stages into surge.
 After shutdown, ammonia refrigeration compressors will reach
settle out . Settle out pressure depends on ambient and loop
volume. This necessitates reducing the loop pressures to
suction pressures for a restart. The seal oil / seal gas systems
are suitably sized for settle out conditions .
 It is generally an expensive proposition to vent off NH3. The
driver (motors) if sized suitably for DOL startup can start the
machine on load. Gas turbine power availability is verified at
lower speeds to ensure compressor accelerates smoothly.
Ammonia Refrigeration compressor
TO B 510

To Vent

From
B505

ASV 1
Nitrogen
from To From
B508 B506 B506
 Startup / Loading :: CO2 Compressors
 CO2 is a high density gas and the CO2 compressors have steep
head curve, sensitive to speed and inlet temperature.
 At startup, stage mismatches won‟t allow the intermediate ASV
to be closed due to lower pressure rise at lower speeds. CO2
compressors are accelerated, in such a way that sufficient time
is given to stabilize the pressures in each section minimize
stage mismatches.
 At higher pressures after cooler, the gas becomes excessively
corrosive, necessitating SS internals for HP section and heating
of CO2 expanding into low pressure zones maintained for
shaft end seals.
 On tripping, the gas will be at various pressure zones, which
settles down to higher pressures . Venting is resorted between
the HP and LP stages during startup and through final vent
during coasting down.
CO2 compressor
Startup / Loading :: Syn. Gas Compressors
 Syn Gas is a low density gas and the syn gas compressors have
reasonably flat head curve.
 At startup, stage mismatches won‟t make a big differences in
the inlet volumes at different stages
 Before the machine reaches close to MGS ( i.e.at lower speeds)
there is a likely hood of mismatch at 2nd suction, a manual
bypass valve is placed is 1st discharge This is closed at lower
speeds say 5000 rpm.
 A HIC is placed after 2nd stage. The 2nd stage by pass is kept
open till MGS, These are evaluated case to case and provided.
 Final antisurge valve is closed for loading the machine to
process.
 Like wise during trip the antiurge valve is assisted by a vent
valve to remove large volume of HP gas accumulated in
discharge pipe
Syn Gas compressor
L P Compresor Turbine H.P Compresor

Vent
 Parallel Operation
 When two or more compressors of similar performance
characteristics are taking suction from a common source
and delivering to a common system, these machines have
to be configured for parallel operation and load sharing.
 However identical the machines are designed and
manufactured, they do not behave absolutely identical due
to differences in operating characteristic or system
resistances like piping valves etc. They do not share the
load automatically equally.
 The machine with higher head rising or flow handling
capability will curtail the other machines range. For large
differences, it will be pushed to surge unless some control
mechanism would come into action.
Parallel Operation
Vendor 1 furnishes Compressor A with operating point X
Vendor 2 furnishes Compressor B for same operating point X,
the system will finely share the load equally at this System
Resistance

Compressor B

Compressor A

Conclusion: Though the characteristics are not same they

will have one common operating point defined by the
System resistance or design point
Load sharing example
An example is given below:
When 3 machines are delivering to a common header,
following can happen.
 A. Two machines may share 50% each and the third
machine may be a hot standby
 B. All the three machines may share 33 .3 % each in
recycle mode
 C. One machine may run at an optimum point of 80%
and the remaining by second machine in recycle mode
and third will remain hot standby or shutdown.
 D. If the load is increased to 150% , then it may be
economical to run 2 machines at 75% each or 100%
first one + 50% second one.
Common Operational Problems
 Vibrations
 Performance deficit
 Bearing temperature rise
 Increased Thrust
 Coupling failure
 Oil Seal /DGS Failures
 Surging of machines
 Auxiliary equipment malfunction
 Oil entry / Gas leakage
 Instrument malfunction
Operational Problems - Vibrations
CAUSE EFFECT

a) Unbalance, misalignment, looseness

b) Higher bearing clearances
c) Operating close to critical speeds
d) Scoring on journals/Thrust collar
e) Worn out couplings Vibrations
f) Surging of machines
g) Improper lubrication
h) External piping forces
I) Flow induced excitations
 Common causes of vibrations
Sl.
Cause Amplitude Frequency Remarks
No.
Increase of vibration
Proportional
Amplitude with speed
1 Unbalance to I X RPM
Common cause of excess
unbalance,
Vibration in machinery
Usually
Best found by appearance of
Misalignmen I X RPM,
Largest in High axial vibrations.
t of often
2 axial Use dial indicators or other
couplings or 2xRPM,
direction methods for positive
bent shaft some-times
3xRPM diagnosis.
Sub-
Journal
harmonics
bearing Primarily Looseness may only develop
3 of RPM,
loose in radial at operating speed
Exactly 1/2
housing
or 1/3 RPM
Common causes of vibrations
Sl.
Cause Amplitude Frequency Remarks
No
Oil whip Machine performance can be
Primarily 42% to 48% Of
4 in journal improved by varying lube oil
radial shaft speed
bearings pressures & temps
Vibrations excited when
Hysteresis Primarily Shaft critical passing thro' critical speeds
5
whirl radial speed and are maintained at higher
speeds.
Damaged Tooth meshing Radial for spur gears, axial
Radial and
6 or wrong frequencies & for helical or herring born
axial
gears harmonics gears
Blade passing
Increased Radial and frequencies Rare cause of trouble except
7
clearances axial No. of vanes x in cases of resonance
RPM
Common causes of vibrations
Sl.
Cause Amplitude Frequency Remarks
No.
At higher speeds vibration
Bearing
8 Radial I x RPM amplitudes increases with
clearance
excessive clearance
High
Looseness of
9 Radial frequency
stator parts
(over 1kHz)
The vibrations will appear at
a particular load and speed
18–50 %
Aero- depending on the source of
50-60% of
10 dynamic Radial instability Such as rotating
running
instability stall in impeller, diffusers or
speed
piping blockage and other
types of flow disturbances
Condition monitoring
 Changes in vibration levels measured on the surface of a
machine are the result of changes in internal forces.
vibration levels increase when the condition of the
machine deteriorates – due to unbalance, misalignment,
bearing & gear tooth wear or due to resonances,
structural modifications, foundation cracks etc.
 Checking machine vibration levels over a period of time
as done in on-line monitoring systems, will indicate the
development of fault conditions This process is called
Machine Condition Monitoring.
Machine Diagnostics
 By using various diagnostic techniques, which also use
vibration measurement as an indicator, the root cause of
the deteriorating machine condition can be established.
 Diagnostic techniques are extremely effective because
they use the information contained in the machine's
vibration signature obtained by frequency analysis of
the machine. It enables troubleshooting of rotor dynamic
problems, rotating component deterioration as well as
structural problems. This process is Machine Diagnostics.
Operational Problems – Performance deficit
CAUSE EFFECT

a) Increase in seal clearances

b) Damaged „O‟ rings Performance
c) Blockage at suction strainer deterioration
d) Blockage of impeller gas Passages
e) Varied suction operating conditions
f) Malfunctioning of valves
(suction /discharge/ASV)
Performance deficit

 Drop in throughput
 Drop in discharge pressure
 Increase in power consumption
 Drop in machine speed
 Increase in discharge temperature
 Unable to load the machine
Operational Problems – Journal Bearings

CAUSE EFFECT

a) Low bearing oil pressures

b) High oil inlet temperatures
c) Unclean oil High bearing
d) Wear out of pads Temperatures
e) Scoring of journal / misalignment
f) Bypassing of oil flow to drain
 Operational Problems – Thrust Bearing

CAUSE EFFECT

a) Low bearing oil pressures

b) High oil inlet temperatures
c) Increased gas thrust High bearing temp.
d) Wear out of pads High axial thrust.
e) Balance drum seal wear out High axial Displacement
f) Inlet oil Bypass to drain
g) Thrust balance gas line Block
h) Wear out of gear teeth
Operational Problems – Gear couplings

CAUSE EFFECT

a) Low lube oil pressures

b) Contaminated oil
c) Misalignment Gear tooth breakage
d) Wear out
e) Excessive torque
f) Unbalance
Operational Problems – Seal oil system
CAUSE EFFECT

a) Low DP between seal oil & ref gas lines

b) Damaged seals/springs/ „o‟ rings
c) Seal oil contamination oil carry over
d) Increased seal clearances or
e) Orifice (vent line) diameter very small Gas leakage
f) Seal oil traps not working
g) LCV/DPCV not working
Operational Problems – Dry gas Seal System
CAUSE EFFECT

a) Primary process gas flow low Primary seal damage

b) Clogged filters Entry of debris Primary seal damage
Machine sealing at risk
c) Secondary N2 pressure low Oil entry into seals.
Gas leakage to atm
d) Improper bearing oil drain Sec. seal damage
e) Condensation of hydrocarbons Primary seal damage
f) Zero leakage from primary/sec. Gas leakage to atm
g) Instrument malfunctioning Machine sealing at risk
 Compressor – Case Study 1
Problem Service Cause / Solution
Thrust CO2 Thrust balancing pipe was too small
bearing which resulted in reduced P
failure across the balancing drum.
Pipe size increased
Compressor – Case Study 2
Problem Service Cause / Solution
Seal rings H2 Oil was contaminated with VGO &
damage recycle block the inlet seal pipes/ports
due to drop resulting in reduced  P.
in P across Oil change, system clean up
seal oil & prevention of entry of VGO into the
Ref. Gas compresso next startup.
Passivate the entire loop with N2
during shutdown and has eliminated
the entry of O2 into the loop
enured prevention of recurrence of
the  P problem.
 Compressor – Case Study 3
Problem Site Cause / Solution
High vibrations at CO2 High pipe line vibrations are
4th Discharge flow induced. Rotating stall
pipe. caused by high gas density
high discharge pressure.
Resolved by modified
diffuser passage in 4th phase.
 Compressor – Case Study 3

REDUCED DIFFUSER INITIAL DIFFUSER

CHANNEL CHANNEL
AFTER BEFORE
 Compressor – Case Study 4

Problem Site Cause / Solution

Frequent tripping Coker Motor was tripping due to
of Coker gas gas high amperage due to higher
mass flow handled by m/c
compressor
because of high suction
during start up pressure. Inlet pressure is not
controllable from process
side.
Suction butterfly valve
replaced with control valve to
drop suction pressure
Compressor – Case Study 4
 Compressor – Case Study 5

Problem Site Cause / Solution

Drop in capacity Wet gas Due to blockage of suction
of wet gas strainer, impeller and
Compressor diaphragm passages with
black dust, the machine
capacity and discharge
pressures dropped.
Machine was cleaned,
overhauled to restore its
normal capacity.
Performance curve with blocked passages

X Normal operating point

Y Operating point with Blocked passages

X
Pd

X’

O
Inlet volumetric flow
 Compressor – Case Study 6
Problem Service Cause / Solution
Surging of CO2 The Compressor was run with
Compressor higher load and reduced
during tripping pressure drops between 2nd &
on load 3rd stages. This caused severe
mismatch and resulted in
surging of LP machine during
trip.
Modified vent valve opening &
recommend to install additional
vent to resolve the surging
problem.
 Compressor – Case Study 6
A to B : DP across Optional process 3 bar : short cut after Optional process removed

7.0 bar 80 bar

25 bar 22 bar 160 bar

1.0 bar

Vent ASV
B
A B

Optional process between A &B

Compressor – Case Study 7
Problem Site Cause / Solution
Improper CO2 Seals frequently getting damaged
Functioning during initial commissioning period.
of Dry Gas The failure was analysed to be due to
Seals. entry of liquid/dirt/corrosion into the
seal.
Clean gas injection line is connected by
throttling from 3rd discharge instead of
from 4th discharge. This prevented
liquid entry into seal.
Compressor – Case Study 7

4th stage 3rd stage

Existing connection Modified connection

 Compressor - Case Study 8

Problem Service Cause / Solution

Low LEF The compressor was experiencing
frequency discharge resistance and tripping on
Vibrations sub-synchronous vibrations during
in H2 start up.
compressor This was identified as a phenomenon
of Rotating stall at 1st phase discharge.
Modification to diffuser area, provision
of shunt holes, increasing bearing size,
were incorporated to reduce the rotor-
dynamic excitation and instability to
eliminate the low frequency vibrations
Compressor - Case Study 8

Low frequency Vibrations- compressor schematic

Compressor - Case Study 8
 Compressor – Case Study 9
Problem Cause / Solution

LPG The compressor tripped due to power failure. And Following

Compressor happened as a consequence
Damage as a 1.Machine tripped due to GT trip for power Failure
result of a trip 2.Vent valve located in suction opened and started venting
3. AOP did not start due to alternate power not being there
4. EOP was not ready and no OH lube oil tank provided
Machine reached 0 speed and continued to run in reverse
Direction due to venting from wrong location.
5.As there was no lube oil supply from any source all the 4
bearings of GT, 3 bearings of compressor were burnt.
Damaging all rotors. Seal system did not suffer major damage
as it had separate SO system..
The summary of errors acted simultaneously are summed up
and corrective actions taken.
Compressor - Case Study 9
 Compressor – Case Study 9
Cause / Solution
The summary of errors acted simultaneously are summed up
and Corrective actions taken.

1. Providing vent valve in suction is prone to cause reverse

direction
2. Not providing a reliable power supply to AOP
3. EOP not provided with back up DC Power
4. OH tank isa good assurance for a coast down situation for
a turbine driven unit as it is likely to take longer time to
coast down.
Damage 7 bearings 3 rotors & downtime of 1 month
Data required for performance evaluation
Inlet /outlet parameters at every phase
 Qs Suction flow in Nm3/hr or Kg/hr
 Ps Suction pressure kg/cm2
 Ts Suction temperature Deg C
 Pd Discharge pressure kg/cm2
 Td Discharge temperature Deg C
 Gas composition to ascertain the molecular weight.
 N Speed rpm
Performance Evaluation
 While flow measurements are made ensure that the
bypass / anti-surge Valves are closed. It is a good
practice to measure the differential pressure across
Orifice and compute the flow.
 Ensure that the respective pressure and temperature
measurements are made with calibrated gauges.
 For ideal or perfect gases the sheet at next slide
gives you the method of evaluating pol. efficiency
and total power consumed
 From the performance curve please mark a point and
see where the operating point lies ?
Parameter Unit Parameter formula
Ps Kg /cm2 abs. Pressure Ratio ρ Pd/Ps

Ts Deg C Temperature Ratio

Ks From gas ( n-1 )/ n

Kd properties
Zs From gas proprties Polytropic efficiency
Zd
Pd Kg /cm2 abs.
Td Deg C Head
Eff.
Z avg (Zs + Zd ) / 2 Flow kg/hr

K avg (Ks +Kd )/ 2 Therm. Power NT kW G * Heff / 102

Flow Qn Nm3/Hr Power loss Mech. NL kW

Mol. wt Mech. losses NM kW

M3/hr Total power NTOTAL kW NT +NL +NM
Performance Evaluation
 Test Power converted to guaranteed conditions
subscript t denote test and g denotes guaranteed


 Where Z =compressibility factor

 M is molecular weight
 T suction temperature
 ρ is pressure ratio
 n polytrophic index
 G mass flow kg/sec
Performance Evaluation
 Corrected Pressure Ratio for guaranteed conditions

 Where Z =compressibility factor

 M is molecular weight
 T suction temperature
 ρ is pressure ratio
 n polytrophic index