fundamentals of turbo ex panders
TRANSCRIPT
Presentation.RPT.DOC
FUNDAMENTALS OF THE
TURBOEXPANDER
“BASIC THEORY AND DESIGN”
PRESENTED BY: MR. JAMES SIMMS
Edited: 3/23/09
i
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
1
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.
2
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.
3
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
4
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.
5
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. These have
been designated Δh's and Δho, respectively.
The ratio of these is the definition of the Isentropic
efficiency, ηe, of the Expander (i.e., ηe = Δho / Δh's ). For
example, if Δho = 34 btu/lb, and Δh's = 40 btu/lb, then:
ηe = 34 / 40 = .85 or 85%
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With the Δh's and ηe values known, we then only need the gas mass
flow rate, w, to calculate the Horsepower developed by the
Expander. Since the mass flow rate is normally given 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 = 20 lbs/second. So, if
Δh's = 40 btu/lb, w = 20 lbs/second, and ηe = .85, 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 - HPBearings
= 961.9 - 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', or Head Rise, developed by the Compressor and its
efficiency, ηc, 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. The
formula is:
HPCompressor = 1.4145 x Δh'Adiabatic x w ___________________________________
ηc
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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. The
units are btu/lb of gas flow. w is the symbol for mass flow
through the Compressor, and the units are lbs/second.
The mass flow rate, w, for the Compressor, is usually given by
the process engineer. 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, ηc, on the
Compressor, such as 75%. Figure 4 on the next page shows a
typical Compression Process in which the ideal head rise, Δh', is
lower than the actual Head Rise, Δho, required to achieve the
Discharge Pressure, the ratio of Δh' / Δho = efficiency of the
Compressor, or ηc.
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We will assume a mass flow rate, w, 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 = Pin x [1 + Δh' ]γ / γ-1
CpT1Z
= 145 x [1 + 18.30 ]1.3 / .3
.54 x (60 + 459.69) x .98
= 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, of the
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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Assume, ACFS2 = 25.0 (ACFS2 is the Actual Cubic Feet per Second
of volumetric flow at the outlet of the Expander). Rearrange
the formula and calculate the speed, 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 U/Co should equal about
0.7. The Co 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 √Δh's _________________ __ = √ 2 x 32.2 x 778 x √40 __________________ __
= √ 50103.2 x √40 __ = 223.8 x √40
= 1415 ft/sec
Since the term U, Wheel Tip Speed, needs to be 0.7 of Co, then:
U = .7 x Co
= .7 x 1415
= 990 ft/sec
DEFINITION OF TERMS:
U = Tip Speed of Wheel, ft/sec
Co = Spouting Velocity, ft/sec
g = 32.2 ft/sec2
J = 778 ft-lb/btu
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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 Δh'Adiabatic to
equal 18.30 btu/lb. Let the head coefficient, Ψ, be equal to .4
and rearrange the formula to solve for U:
U2 = Δh' x g x J
.4 ______________________
U = √(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.
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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. The oil is then filtered by one of the Dual
Filters. 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.
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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.
20
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.
21
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.
22
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. In
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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DESCRIPTION OF SURGE CONTROL SYSTEM: THE SURGE CONDITION:
The phenomenon of “surge” in an axial or centrifugal Compressor
occurs when the flow is reduced sufficiently to cause a
momentary reversal of flow. This reversal tends to lower the
pressure in the discharge line. Normal compression resumes, and
the cycle is repeated. This cycling, or surging, can vary in
intensity from an audible rattle to a violent shock. Intense
surges are capable of causing complete destruction of the
components in the Compressor such as Blades, Bearings, and
Seals. 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. The function
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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. In operation, the Inlet
Flow is 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
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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.
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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 , 50 + 130 = 180 PSIG 3
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.
29
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:
________________________________________________________________
________________________________________________________________
________________________________________________________________
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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. 2. CRACKED HEAT BARRIER, SYMPTOMS SAME AS ABOVE.
1. DISASSEMBLE UNIT, INSPECT AND CHANGE HEAT BARRIER, OR EXPANDER SEAL INSERT AND EXPANDER SHAFT SEAL RING IF NECESSARY. 2. SAME AS ABOVE.
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.
32
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. OIL BYPASSING COOLER. 2. CONDENSATE MIXING WITH OIL RESULTING IN LOWER VISCOSITY. 3. EXCESSIVE THRUST UNBALANCE.
1. CHECK TO SEE IF OIL CONTROL VALVE IS OPEN. 2. CHECK OIL LEVEL, CHANGE IF NECESSARY. 3. THIS CONDITION MAY WEAR THE JOURNAL BEARINGS RESULTING IN A HIGHER RUNNING TEMPERATURE. DISASSEMBLE, INSPECT, & REPAIR.
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.
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.
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 DE-METHANIZER TO THE BOOSTER COMPRESSOR SUCTION.
1. SEE - HIGH COMPRESSOR BEARING TEMPERATURE OR OIL DRAIN TEMPERATURE (SET POINT SPECIFICATIONS) 2. CHECK FOR BLOCKAGE OR ABNORMAL PRESSURE DROP IN HEAT EXCHANGER.
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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.