FUNDAMENTALS OF THE TURBOEXPANDER “BASIC THEORY AND DESIGN”
PRESENTED BY: MR. JAMES SIMMS
Edited: 3/23/09
Presentation.RPT.DOC
TABLE OF CONTENTS
INTRODUCTION/DESCRIPTION .................................... A APPLICATION ................................................. B PRELIMINARY SIZE CALCULATIONS ............................... C ENERGY EXTRACTION ........................................... D HORSEPOWER BALANCE .......................................... E HOW DOES THE BOOSTER COMPRESSOR USE THE HORSEPOWER TO MAKE PRESSURE? ................................................... F DISCHARGE PRESSURE CALCULATION .............................. G HOW IS THE TURBOEXPANDER DESIGN SPEED DETERMINED? ........... H HOW TO APPROXIMATE THE EXPANDER WHEEL DIAMETER .............. I HOW TO APPROXIMATE THE COMPRESSOR WHEEL DIAMETER ............ J TURBOEXPANDER CONTROLS, WITH REFERENCE TO THE TURBOEXPANDER SYSTEM SCHEMATIC DWG (FIGURE 5) ............................. K LUBE OIL SYSTEM ............................................. K SHAFT SEALING SYSTEM ........................................ L DESCRIPTION OF THRUST BALANCE SYSTEM ........................ M DESCRIPTION OF SURGE CONTROL SYSTEM ......................... N MAINTENANCE ................................................. O IMPORTANCE OF LOGGING DATA .................................. P EXAMPLE OF DATA LOG SHEET ................................... P TROUBLE SHOOTING ............................................ Q ABOUT THE AUTHOR ............................................ R
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FUNDAMENTALS OF THE TURBOEXPANDER BASIC THEORY & DESIGN
INTRODUCTION/DESCRIPTION: The term "Turboexpander", Figure 1, is normally used to define an Expander/Compressor machine as a single unit.
It consists of
two (2) primary components; the Radial Inflow Expansion Turbine and a Centrifugal (Booster) Compressor combined as an assembly. Its Wheels are connected on a single Shaft.
The Expansion
Turbine is the power unit and the Compressor is the driven unit.
1
In a Gas Processing Plant, the purpose of the Turboexpander is to efficiently perform two (2) distinctly different, but complimentary, functions in a single machine.
The primary
function is to efficiently generate refrigeration in the process gas stream.
This is done by the Expansion Turbine end
efficiently extracting the potential heat energy from the gas stream, causing it to cool dramatically.
This extracted energy
is converted to mechanical energy to rotate the Shaft to the Booster Compressor end of the Turboexpander, which partially recompresses the residue gas stream.
The Turboexpander operates
according to the thermodynamic and aerodynamic laws of physics. When designed properly, the Turboexpander can yield very high efficiencies at the "Design Point" and reasonable efficiencies at other, or "Off-Design", Points.
APPLICATION: The typical Turboexpander process installation is shown in Figure 2, Simplified Process Schematic.
High pressure,
moderately cold gas flows into the Expander section of the Turboexpander.
The gas flows through the Expander Variable
Inlet Nozzles (Guide Vanes) and then through the Wheel, exhausting at a lower pressure and at a substantially colder temperature.
Gas flows from the Expander to the Demethanizer,
where condensate is removed. 2
It should be noted that the Expander Nozzles are used to control the gas flow rate in order to maintain the pressure in the Demethanizer.
The Residue Gas from the Demethanizer Tower flows
through the Feed Gas Heat Exchanger and then to the Booster Compressor end of the Turboexpander.
The efficiency of the
Booster Compressor is very important, as it can improve the expansion process for more refrigeration as well as more efficiently use the power extracted by the Expander.
While the
engineering process to properly design a high efficiency Turboexpander is very complex, requiring computerized analytical 3
tools, the basic initial sizing process can be simplified by using certain basic equations and assumptions.
PRELIMINARY SIZE CALCULATIONS: The Turboexpander operation is best described as a dynamic system which responds to the Process Stream variations.
To
start the initial sizing design process, a fixed set of Process Stream parameters, or "Design Point", must be established.
This
"Design Point" is normally set by the plant process engineer's system analysis in the case of new plants.
In the case of a
Turboexpander re-design in an existing plant, the actual operating condition will dictate the new "Design Point". The parameters required to size the Turboexpander are:
Gas Composition
Flow Rate
Inlet Pressure
Inlet Temperature
Normally, the Expander Outlet Pressure is determined by the performance of the Booster Compressor's efficiency through a complex iterative analysis.
For this simplified sizing
exercise, we will assume the value of the Expander Outlet Pressure. 4
ENERGY EXTRACTION: Where does the energy come from?
With reference to Figure 3,
which shows the typical Expansion Process Graph, we see the process in terms of change of energy, or
Δh,
with units of
btu/lb, both in the Ideal Process (Isentropic, or 100% efficient) and in the Actual Process, or real terms. been designated
Δh's and Δho,
These have
respectively.
The ratio of these is the definition of the Isentropic efficiency,
ηe,
example, if
Δho = 34 btu/lb,
of the Expander (i.e., and
ηe = Δho / Δh's ).
Δh's = 40 btu/lb,
ηe = 34 / 40 = .85 or 85% 5
then:
For
With the
Δh's
and
ηe
values known, we then only need the gas mass
flow rate,
w,
Expander.
Since the mass flow rate is normally given by the
to calculate the Horsepower developed by the
process engineer, we now have enough information to calculate the Horsepower with the following formula:
HPExpander = (778 / 550) x Δh's x w x ηe 778 / 550 = 1.4145 (This is the constant value to change btu units into Horsepower terms) We will assume the mass flow rate,
w
Δh's = 40 btu/lb, w = 20 lbs/second,
and
= 20 lbs/second.
ηe = .85,
So, if
then:
HPExpander = 1.4145 x 40 x 20 x .85 = 961.9
HORSEPOWER BALANCE: The HP developed by the Expander must be absorbed in order to prevent over-speeding.
The Bearings and the Compressor absorb
this power to create a balance.
The Horsepower Balance formula
is:
HPExpander = HPCompressor + HPBearings
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Example: Let's assume that the Bearings consume a total of 30 Horsepower, which is subtracted from the Expander Horsepower.
If we
rearrange the formula above, then the available Horsepower to the Compressor is:
HPCompressor
=
HPExpander
=
961.9
- HPBearings -
30
= 931.9 Horsepower
HOW DOES THE BOOSTER COMPRESSOR USE THE HORSEPOWER TO MAKE PRESSURE?: The pressure rise through the Compressor is a function of the adiabatic
Δh',
efficiency,
ηc,
or Head Rise, developed by the Compressor and its to satisfy the Horsepower balance formula.
The
Horsepower formula for the Compressor is similar to the Expander, but slightly different as it is consuming power. formula is:
HPCompressor = 1.4145 x Δh'Adiabatic x w
___________________________________
ηc 7
The
The value of 1.4145 is, again, the constant to convert the units into Horsepower, just the same as the Expander formula.
Δh'Adiabatic
is the ideal Head Rise, or energy change, from the
Inlet Pressure to the Discharge Pressure of the Compressor. units are btu/lb of gas flow.
w
The
is the symbol for mass flow
through the Compressor, and the units are lbs/second.
The mass flow rate,
w,
the process engineer.
for the Compressor, is usually given by With the available Compressor Horsepower
known, we can rearrange the Compressor Horsepower formula to calculate the
Δh'Adiabatic,
Head Rise, as follows:
Δh'Adiabatic = HPCompressor x ηc
______________________
1.4145 x w
We start by assuming a reasonable efficiency, Compressor, such as 75%.
ηc,
on the
Figure 4 on the next page shows a
typical Compression Process in which the ideal head rise, lower than the actual Head Rise,
Δho,
Discharge Pressure, the ratio of
Δh' / Δho
Compressor, or
ηc. 8
Δh',
required to achieve the = efficiency of the
is
w,
We will assume a mass flow rate,
through the Compressor, of
27.0 lbs/second and solve the formula:
Δh'Adiabatic =
931.9 x .75 1.4145 x 27.0
= 18.30 btu/lb
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DISCHARGE PRESSURE CALCULATION: Now, having calculated the
Δh'Adiabatic,
we can calculate the
pressure rise through the Compressor by the following formula:
Pout = Pin x [1 +
Δh' ]γ / γ-1 CpT1Z
DEFINITION OF TERMS:
Pout
= Discharge Pressure, PSIA
Pin
= Inlet Pressure, PSIA
Cp
= Specific heat at constant pressure
T1
= Inlet temperature, °R (°R = °F + 459.69)
Z
= Compressibility Factor
γ
= Greek Symbol "Gamma", stands for the Ratio of specific heats
To be accurate, the values of
Cp, Z,
and
γ
are best derived by a
good equation of state computer program; however, for this exercise, we will assume certain values for these (T1 = 60°F,
Cp
= .54,
Z
= .98, and
γ
= 1.3) and solve the above formula.
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So, to repeat the formula for the Compressor, if the Inlet Pressure equals 145 PSIA, the Discharge Pressure will be calculated:
Pout
=
Δh' CpT1Z
Pin x [1 +
]γ / γ-1
18.30 ]1.3 / .3 .54 x (60 + 459.69) x .98
= 145 x [1 +
= 145 x [ 1.0665 ] 4.333 =
145 x 1.322
=
191.7 PSIA
HOW IS THE TURBOEXPANDER DESIGN SPEED DETERMINED?: This is done by using the term Specific Speed, or
Ns,
Expander Wheel:
Ns
=
______ N x √ACFS2 ________________________
( 778 x Δh's ) .75 From earlier pages,
Δh's = 40 btu/lb
Assume a Specific Speed,
Ns = 75
for good efficiency
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of the
Assume,
ACFS2 = 25.0 (ACFS2
is the Actual Cubic Feet per Second
of volumetric flow at the outlet of the Expander). the formula and calculate the speed,
Rearrange
N:
N = Ns x ( 778 x Δh's ) .75 ____________________ ______ √ ACFS2 = 75 x (778 x 40) .75 __ √25 = 75 x 2343 5 = 35,146 RPM
HOW TO APPROXIMATE THE EXPANDER WHEEL DIAMETER: First, for good efficiency, the term
0.7.
The
Co
U/Co
should equal about
term is known as the spouting velocity of the gas
from the Nozzles into the Expander Wheel.
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The formula is:
_____________
Co =
√2 x g x J
x _________________ = √ 2 x 32.2 x 778
___
√Δh's __ x √40
__________________
=
Since the term
U,
√
x
=
223.8
__ x √40
=
1415 ft/sec
Wheel Tip Speed, needs to be
= .7 x 1415 = 990 ft/sec
DEFINITION OF TERMS:
Co
= =
Tip Speed of Wheel, ft/sec Spouting Velocity, ft/sec
g
=
32.2 ft/sec2
J
=
778 ft-lb/btu
√40
50103.2
U = .7 x Co
U
__
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0.7
of
Co,
then:
The term,
U,
is the peripheral velocity of the Expander Wheel
(tip speed) which is calculated by:
U = Diameter (inches) x RPM 229.2
So, to rearrange the formula to solve for the Wheel Diameter:
Diameter =
U x 229.2 RPM
=
990 x 229.2 35,146
= 6.46 inches
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HOW TO APPROXIMATE THE COMPRESSOR WHEEL DIAMETER: The following formula can be used to approximate the Compressor Wheel Diameter:
Δh'Adiabatic = U2 x Ψ
____________
g x J From previous pages, we calculated the Compressor equal 18.30 btu/lb.
Let the head coefficient,
and rearrange the formula to solve for
U2
=
U =
Ψ,
Δh'Adiabatic
to
be equal to .4
U:
Δh' x g x J .4 ______________________
√(Δh'
x g x J) / .4
_____________________________ U =
√(18.30
x 32.2 x 778) / .4
____________ U =
√1,146,110.7
U = 1070.6 ft/sec
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Now calculate the Compressor Wheel Diameter from the equation:
U = Wheel Diameter x RPM 229.2
From the Expander analysis, the design speed was calculated to be 35,146 RPM.
So rearranging and solving the above formula:
Compressor Wheel diameter =
U
x 229.2 RPM
= 1070.6 x 229.2 35,146 = 6.98 inches ** CAUTION NOTE: While the above method does give the steps necessary to do the initial rough sizing of the Expander and Compressor Wheels, it is vastly over simplified.
In the above, certain values were
assumed for specific speed, efficiency, head coefficient, etc. These assumptions were used for ease of demonstrating the formulas only.
These assumptions MUST NOT be used as "rule of
thumb" values, as they may give false results in an actual case. Any actual design analysis should be performed by an experienced Turboexpander Design Engineer.
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TURBOEXPANDER CONTROLS, WITH REFERENCE TO THE TURBOEXPANDER SYSTEM SCHEMATIC DWG (FIGURE 5): The Turboexpander system pressure basically floats with the Compressor suction pressure.
This is achieved by venting the
pressurized oil Reservoir through a De-misting Pad back to the Compressor Suction.
LUBE OIL SYSTEM: The Lube Oil Pressure supplied to the Bearings is controlled at a pressure higher than the Reservoir, typically this pressure is 150 PSID controlled by a Differential Pressure Regulating Valve relieving excess oil to the Reservoir. 17
Dual, Electric Motor Driven, Lube Oil Pumps (one Main and one Standby) take suction from the Reservoir providing oil through the Cooler or bypassing the Cooler via the Temperature Control Valve as required to maintain a preset oil supply temperature to the Bearings. Filters.
The oil is then filtered by one of the Dual
At this point the oil pressure regulates to the proper
Differential Pressure.
An Accumulator is installed to store oil
for emergency coast down in case of electrical power outage. The oil then travels to the Bearings.
Some systems have an oil
flow rate measuring device in the oil supply line to the Bearings.
SHAFT SEALING SYSTEM: The Shaft Seals are Labyrinth type using Seal (Buffer) Gas to prevent cold gas migration into the Bearing Housing and to prevent oil leakage into the process stream.
The Seal Gas system typically uses warm process gas that has been filtered.
The pressure to the Labyrinth Seals is
maintained at a Differential Pressure of typical 50 PSID above the pressure behind the Expander Wheel.
This is accomplished by
use of a Differential Pressure Regulator sensing and floating with the Expander Back Wheel Pressure. 18
The small amount of Seal Gas going across the Seal towards the Bearing Housing mixes with the oil and drains to the Reservoir. In the Reservoir, this separates from the oil and then is vented through the De-misting Pad out into the Booster Compressor Suction.
Basically, the Turboexpander System Pressure
automatically floats with the process pressure via the Compressor Suction.
For control of the Turboexpander at start-up, a local Hand Indicating Controller (HIC) is provided to adjust the Expander variable Nozzles to control the Gas Flow (and Expander Speed) during start-up.
Primary safety instrumentation includes Speed Probe, Vibration Probe, Bearing Temperature RTDs, Thrust Oil, and Seal Gas Differential Pressure Switches.
A first-out Annunciator System
provides an indication of the initial cause of a shut-down.
The Automatic Thrust Equalizer System vents pressure from behind the Compressor Wheel to try to equalize the oil pressure measured at each Thrust Bearing.
19
DESCRIPTION OF THRUST BALANCE SYSTEM: The purpose of this system is to control the Axial Thrust of the Expander/Compressor Rotor within safe load limits, in either direction, on the Thrust Bearings.
The original version (Figure 6) operates in the following manner.
The oil pressure at the Thrust Face of each Bearing is
measured and connected to two opposite ends of a Piston Chamber. A Piston Shaft is connected with an Internal Seal to a Gate Valve that is connected between the pressure port behind the Compressor Wheel and the suction of the Compressor.
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The GTS improved version (Figure 7) operates in a similar manner as the original system, but with the addition of a constant feed oil supply for improved response to Thrust variations.
The Gate
Valve is replaced by a "balanced piston" Spool Valve, which is less prone to sticking (less resistance) than the Gate Valve, and therefore more responsive.
The Spool Valve is connected
between the pressure vent port behind the Compressor Wheel and the suction of the Compressor, just as the original system.
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The Thrust balancing is accomplished by maintaining or decreasing the pressure behind the Compressor Wheel.
This is
accomplished when the Thrust force (oil pressure) on the Compressor Thrust Bearing is increased, which causes the Piston to slowly open the Spool Valve, thereby reducing the pressure in the area behind the Compressor Wheel causing the load on the Compressor Thrust Bearing to reduce.
The opposite occurs when
the load is increased toward the Expander Thrust Bearing. turn, the Spool Valve will tend to close.
With the GTS version,
the end of the Thrust Equalizing Valve can be viewed and verified as to which direction the Thrust Valve is in.
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In
DESCRIPTION OF SURGE CONTROL SYSTEM:
THE SURGE CONDITION: The phenomenon of “surge” in an axial or centrifugal Compressor occurs
when
the
flow
is
reduced
momentary reversal of flow.
to
cause
a
This reversal tends to lower the
pressure in the discharge line. the cycle is repeated.
sufficiently
Normal compression resumes, and
This cycling, or surging, can vary in
intensity from an audible rattle to a violent shock.
Intense
surges
of
the
Bearings,
and
are
components Seals.
capable in
the
of
causing
Compressor
complete
such
as
destruction
Blades,
An Anti-Surge Control system is therefore recommended to
provide positive protection against surging or cycling.
FLOW RELATION IN THE COMPRESSOR: The performance map of the Compressor, typically supplied by the Compressor manufacturer, is a plot of Pressure increase (head) vs. Capacity (flow) over the full range of operating conditions. At any given Compressor speed, a point of maximum discharge pressure is reached as the flow is reduced.
This is indicated
at points A, B, C, D, E, and F in Figure 8 on the next page. The line connecting these points describes the surge limit line, which separates the region of safe operation from the surge area.
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As flow is reduced, the pressure in the Compressor tends to be lower than the discharge pressure and a momentary reversal of flow occurs.
This reversal then tends to lower the pressure at
the discharge, allowing for normal compression until the cycle repeats itself.
- FIGURE 8 -
SURGE CONTROL: The most common method of Surge Control uses the Compressor ∆P to represent “head” and the differential pressure across an Inlet orifice (called “h”) to represent capacity.
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The function
of the Surge Control system is to keep the ratio of
∆P/h
from
exceeding the slope of the surge line. To provide some factor of safety, a control line(Set Point) should be established to the right of the surge line, as shown in Figure 9 below.
- FIGURE 9 -
A typical anti-Surge control system block diagram is shown on the sketch on the next page, Figure 10. Flow is
In operation, the Inlet
measured differentially through an Orifice plate using
a DP Transmitter, with its output connected to a Ratio Station (multiplier) and then transmitted as a set point to the Controller(PID).
The Inlet and Discharge pressures of the
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Compressor are also measured using a DP Transmitter and transmitted as the Process variable to the Controller.
The output of the Controller operates a bypass valve that recycles the gas back to the Inlet of the Compressor.
In
general, when the two (2) signals become equal within the bounds of the Proportional Band range, the Controller will output a signal to open the Bypass (Recirculation) Valve.
If the Flow
rate is too low (i.e., near surge) then the Bypass Valve should open and allow additional flow to re-circulate back to the 26
Compressor as necessary.
In normal operation, the Bypass Valve
is closed to prevent pressure losses and it opens only to prevent surge.
MAINTENANCE: The general rule is that there is no need for a strict, regular maintenance schedule on the Turboexpander.
There are no parts
that are "wear parts" or parts that have a finite life expectancy.
The exception to this general rule is when there is
some known issue that may demand regular maintenance surveillance.
For example, wet Inlet Process Gas that may cause
erosion on the Expander Nozzles or Wheel.
The main component on the Turboexpander System that should be checked on a regular basis is the pre-charge pressure in the Lube Oil Accumulator's Bladder.
This is important because the
Accumulator pre-charge pressure is critical to providing emergency oil to the Bearings for a safe coast down during an electrical power outage.
This pre-charge pressure can only be
checked when there is no oil supply pressure to the Accumulator (for example, when the pumps are stopped).
If the Accumulator
must be checked while the pumps are running, then it must be isolated from the oil line by shutting off the Block Valve. 27
To accurately check the pre-charge pressure, the oil in the Accumulator must be drained below the Bladder pre-charge pressure.
The Bladder should be pre-charged using N2 gas to
pressure of above ⅓ the oil differential pressure plus the normal Reservoir pressure.
For example, DP = 150 PSID and the Reservoir normal operating pressure is 130 PSIG, therefore:
150 = 50 , 3
50 + 130 = 180 PSIG
So the pre-charge pressure should be above 180 PSIG.
This is an
approximate value, so the tolerance can be ± 5 PSIG without creating a problem.
Changing the Lube Oil in the reservoir is normally not necessary unless there is a cause to believe that it is contaminated or its viscosity has changed.
Normally a periodic sampling of the
oil to check viscosity is recommended.
Perhaps twice a year is
sufficient unless some event occurs to suggest a more frequent check is needed.
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There have been several installations that have process streams with unusually high molecular weight due to heavy hydrocarbons. The installation did experience some dilution of the oil which reduced the viscosity.
The solution was to mix higher viscosity
oil of the same type to achieve the desired viscosity.
In these
installations, the oil viscosity is rechecked frequently, perhaps monthly.
IMPORTANCE OF LOGGING DATA: A very important trouble-shooting tool is reviewing the operational history of the Turboexpander.
The best warning sign
of a possible impending problem is a change in the operational characteristics of the Turboexpander, i.e. Bearing temperature, vibration, and sound.
For this reason, it is very important to
maintain good data log sheets.
Shown on the following sheet is
the list of data points to be regularly recorded on data log sheets.
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EXAMPLE OF DATA LOG SHEET *DATE/TIME *FLOW RATE *SIGNAL TO EXPANDER ACTUATOR *EXPANDER INLET PRESSURE *EXPANDER INLET TEMPERATURE *EXPANDER OUTLET PRESSURE *EXPANDER OUTLET TEMPERATURE *COMPRESSOR INLET PRESSURE *COMPRESSOR INLET TEMPERATURE *COMPRESSOR OUTLET PRESSURE *COMPRESSOR OUTLET TEMPERATURE *SHAFT SPEED *VIBRATION *EXPANDER BEARING RTD *COMPRESSOR BEARING RTD *EXPANDER THRUST PRESSURE *COMPRESSOR THRUST PRESSURE *SEAL GAS SUPPLY PRESSURE *EXPANDER BACK WHEEL PRESSURE *SEAL GAS FILTER DP *LUBE OIL PRESSURE AT PUMP DISCHARGE *LUBE OIL SUPPLY PRESSURE TO BEARINGS *LUBE OIL RESERVOIR PRESSURE TO BEARINGS *LUBE OIL FLOW INDICATOR *LUBE OIL TEMPERATURE (UPSTREAM OF COOLER) *LUBE OIL TEMPERATURE (DOWNSTREAM OF COOLER) *LUBE OIL FILTER AP *MISCELLANEOUS
OPERATOR COMMENTS: ________________________________________________________________ ________________________________________________________________ ________________________________________________________________ 30
TROUBLE SHOOTING CONDITION:
PROBABLE CAUSES:
REMEDY:
COLD OIL DRAIN TEMPERATURE
LOW SEAL GAS PRESSURE, ALLOWING COLD PROCESS GAS TO ENTER JOURNAL BEARING HOUSING.
INCREASE SEAL GAS PRESSURE. CHECK OIL DRAIN TEMPERATURE FOR INCREASE.
A DROP IN OIL DRAIN TEMPERATURE OF LESS THAN 50°F IS NOT UNCOMMON. HOWEVER, SEAL GAS PRESSURE ADJUSTMENTS WILL GENERALLY REMEDY SUCH A TEMPERATURE CHANGE. IF NOT-
1. THE LABYRINTH SECTION OF THE HEAT BARRIER HAS WASHED OUT. THE SEAL GAS CAN NO LONGER BUFFER THE COLD PROCESS.
1. DISASSEMBLE UNIT, INSPECT AND CHANGE HEAT BARRIER, OR EXPANDER SEAL INSERT AND EXPANDER SHAFT SEAL RING IF NECESSARY.
COMPRESSOR SURGE
NOTE: SURGE WILL OCCUR UNTIL 70% OF DESIGN FLOW IS ATTAINED. THE TACHOMETER WILL FLUCTUATE 150 TO 200 RPM ABOUT EVERY 6 SECONDS. CAUTION: OPERATING UNDER SURGE CONDITIONS FOR MORE THAN 10-15 MINUTES MAY DAMAGE THE BEARINGS.
KEEP THE COMPRESSOR BYPASS OPEN UNTIL THE EXPANDER REACHES 80% OF DESIGN CAPACITY.
HIGH COMPRESSOR TEMPERATURE
COMPRESSOR BYPASS NOT CLOSED, CAUSING RECIRCULATION OF HOT GASES RESULTING IN COMPOUND TEMPERATURE INCREASES.
CLOSE BYPASS. CAUTION: THE COMPRESSOR IS THE LOAD FOR THE EXPANDER. ANY CHANGE IN THE FLOW WILL AFFECT THE OVERALL PROCESS CONDITIONS, SPEED, TEMPERATURE, ETC.
2. CRACKED HEAT BARRIER, SYMPTOMS SAME AS ABOVE.
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2. SAME AS ABOVE.
COMPRESSOR IMPELLER BACK SEAL WASHED OUT
HIGH COMPRESSOR DISCHARGE TEMPERATURE
CHECK PLAUSIBLE CAUSE OF COMPRESSOR HIGH TEMPERATURE, SAME AS ABOVE.
HIGH OIL TEMPERATURE
NOTE: OIL TEMPERATURE MAY VARY WITH CHANGES IN AMBIENT TEMPERATURE.
1. CHECK TO SEE IF OIL CONTROL VALVE IS OPEN.
1. OIL BYPASSING COOLER. 2. CONDENSATE MIXING WITH OIL RESULTING IN LOWER VISCOSITY. 3. EXCESSIVE THRUST UNBALANCE. UNBALANCED THRUST TOWARD THE EXPANDER BEARING.
1. EXPANDER ROTOR BACK SEAL WASHED OUT ALLOWING INLET PRESSURE TO BE FELT BEHIND THE ROTOR LOADING THE EXPANDER THRUST BEARING. 2. EXPANDER ROTOR RELIEF HOLES (IF PROVIDED) PLUGGED, CAUSING PRESSURE BUILD-UP BEHIND THE ROTOR LOADING THE THRUST BEARING.
UNBALANCED THRUST TOWARD COMPRESSOR BEARING
1. WASHED OUT COMPRESSOR IMPELLER BACK SEAL, CAUSED BY EXCESSIVE COMPRESSOR TEMPERATURE. 2. ABNORMAL PRESSURE DROP IN THE RESIDUE GAS LINE FROM THE DEMETHANIZER TO THE BOOSTER COMPRESSOR SUCTION.
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2. CHECK OIL LEVEL, CHANGE IF NECESSARY. 3. THIS CONDITION MAY WEAR THE JOURNAL BEARINGS RESULTING IN A HIGHER RUNNING TEMPERATURE. DISASSEMBLE, INSPECT, & REPAIR. 1. CHECK COMPRESSOR THRUST BALANCER TO INSURE PROPER ADJUSTMENT FOR PARTICULAR RUN CONDITIONS. 2. THE DEHYDRATOR UPSTREAM OF THE EXPANDER IS SATURATED OR INOPERATIVE. THE ROTOR MUST BE THAWED OUT BY A WARM GAS STREAM. 1. SEE - HIGH COMPRESSOR BEARING TEMPERATURE OR OIL DRAIN TEMPERATURE (SET POINT SPECIFICATIONS) 2. CHECK FOR BLOCKAGE OR ABNORMAL PRESSURE DROP IN HEAT EXCHANGER.
FROZEN SEAL GAS LINE
SEAL GAS PRESSURE TOO LOW, PROCESS GAS BACKS UP INTO SEAL GAS LINE.
INSTALL A DIFFERENTIAL PRESSURE REGULATOR IN THE EXPANDER "WHEEL TAP" TO MAINTAIN SEAL GAS PRESSURE ABOVE WHEEL TAP PRESSURE BY 20 TO 50 PSI.
FOR ANY ASSISTANCE OR TROUBLESHOOTING QUESTIONS PLEASE FEEL FREE TO CONTACT OUR ENGINEERING SERVICE DEPARTMENT.
SIMMS MACHINERY INTERNATIONAL, INC. 2357 “A” STREET SANTA MARIA, CALIFORNIA 93455 U.S.A. TEL: 805-349-2540 FAX: 805-349-9959 E-MAIL:
[email protected] WEBSITE: www.simmsmachineryinternational.com
GAS TECHNOLOGY SERVICES, INC. 2357 “A” STREET SANTA MARIA, CALIFORNIA 93455 U.S.A. TEL: 805-349-0258 FAX: 805-349-9959 E-MAIL:
[email protected] WEBSITE: www.gastechnologyservices.com
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ABOUT THE AUTHOR: James (Jim) Simms has been involved with the design, manufacture, and service of Turboexpanders and other Cryogenic Rotating Machinery since 1969.
He worked in the Engineering
Department of various Turboexpander manufacturing companies until he founded Simms Machinery International, Inc. in 1988. The company primarily focuses on LNG Boil-off Gas Compressors aboard LNG Tankers (Ships).
In 1994 he, along with others,
founded Gas Technology Services, Inc. to service Turboexpanders, primarily used in Gas Processing Plants.
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