Pressure Control (Turbo-Propeller Engine)

The pressure control system (Figure 7–4) is a typical system as fitted to a turbo-propeller engine where the rate of engine acceleration is restricted by a propeller speed controller. The fuel pump output is automatically controlled by spill valves in the flow control unit (F.C.U.) and the engine speed governor. These valves, by varying the fuel pump servo pressure, adjust the pump stroke to give the correct fuel flow to the engine.

At steady running conditions, at a given air intake pressure and below governed speed, the spill valve in the F.C.U. is in a sensitive position, creating a balance of forces across the fuel pump servo piston and ensuring a steady pressure to the throttle valve.

When the throttle is slowly opened, the pressure to the throttle valve falls and allows the F.C.U. spill valve to close, so increasing the servo pressure and pump delivery. As the pressure to the throttle is restored, the spill valve returns to its sensitive or controlling position, and the fuel pump stabilizes its output to give the engine speed for the selected throttle position. The reverse sequence occurs as the throttle is closed.

A reduction of air intake pressure, due to a reduction of aircraft forward speed or increase in altitude, causes the F.C.U. capsule to expand, thus increasing the bleed from the F.C.U. spill valve. This reduces fuel pump delivery until the fuel flow matches the airflow and the reduced H.P. pump delivery (throttle inlet pressure), allows the spill valve to return to its sensitive position. Conversely, an increase in air intake pressure reduces the bleed from the spill valve and increases the fuel flow. The compensation for changes in air intake pressure is such that fuel flow cannot be increased beyond the predetermined maximum permissible for static International Standard Atmosphere (I.S.A.) sea-level conditions.

The engine speed governor prevents the engine from exceeding its maximum speed limitation. With increasing engine speed, the centrifugal pressure from the fuel pump rotor radial drillings increases and this is sensed by the engine speed governor diaphragm. When the engine reaches its speed limitation, the diaphragm is deflected to open the governor spill valve, thus overriding the F.C.U., and preventing any further increase in fuel flow. Some pressure control systems employ a hydro-mechanical governor.

The governor spill valve also acts as a safety relief valve. If the fuel pump delivery pressure exceeds its maximum controlling value, the servo pressure acting on the orifice area of the spill valve forces the valve open regardless of the engine speed, so preventing any further increase in fuel delivery pressure.

Pressure Control (Turbo-Jet Engine)

In the pressure control system illustrated in Figure 7–5, the rate of engine acceleration is controlled by a dashpot throttle unit. The unit forms part of the fuel control unit and consists of a servo-operated throttle, which moves in a ported sleeve, and a control valve. The control valve slides freely within the bore of the throttle valve and is linked to the pilot’s throttle by a rack and pinion mechanism. Movement of the throttle lever causes the throttle valve to progressively uncover ports in the sleeve and thus increase the fuel flow. Figure 7–6 shows the throttle valve and control valve in their various controlling positions.

At steady running conditions, the dashpot throttle valve is held in equilibrium by throttle servo pressure opposed by throttle control pressure plus spring force. The pressures across the pressure drop control diaphragm are in balance and the pump servo pressure adjusts the fuel pump to give a constant fuel flow.

When the throttle is opened, the control valve closes the low-pressure (L.P.) fuel port in the sleeve and the throttle servo pressure increases. The throttle valve moves towards the selected throttle position until the L.P. port opens and the pressure balance across the throttle valve is restored. The decreasing fuel pressure difference across the throttle valve is sensed by the pressure drop control diaphragm, which closes the spill valve to increase the pump servo pressure and therefore the pump output. The spill valve moves into the sensitive position, controlling the pump servo mechanism so that the correct fuel flow is maintained for the selected throttle position.

During initial acceleration, fuel control is as described previously; however, at a predetermined throttle position the engine can accept more fuel and at this point the throttle valve uncovers an annulus, so introducing extra fuel at a higher pressure (pump delivery through one restrictor). This extra fuel further increases the throttle servo pressure, which increases the speed of throttle valve travel and the rate of fuel supply to the spray nozzle.

On deceleration, movement of the control valve acts directly on the throttle valve through the servo spring. Control valve movement opens the flow ports through the control valve and throttle valve, to bleed servo fuel through the L.P. port. Throttle control pressure then moves the throttle valve towards the closed position, thus reducing the fuel flow to the spray nozzles.

Changes in air intake pressure, due to a change in aircraft altitude or forward speed, are sensed by the capsule assembly in the fuel control unit. With increased altitude and a corresponding decrease in air intake pressure, the evacuated capsule opens the spill valve, so causing a reduction in pump stroke until the fuel flow matches the airflow. Conversely, an increase in air intake pressure closes the spill valve to increase the fuel flow.

H.P. compressor shaft rpm is governed by a hydro-mechanical governor that uses hydraulic pressure proportional to engine speed as its controlling parameter. A rotating spill valve senses the engine speed and the controlling pressure is used to limit the pump stroke and so prevent overspeeding of the H.P. shaft rotating assembly. The controlling pressure is unaffected by changes in fuel specific gravity.

At low H.P. shaft speeds, the rotating spill valve is held open, but as engine speed increases, centrifugal loading moves the valve towards the closed position against the diaphragm loads. This restricts the bleed of fuel to the L.P. side of the valve until, at governed speed, the governor pressure deflects the servo control diaphragm and opens the servo spill valve to control the fuel flow and thereby the H.P. shaft speed.

If the engine gas temperature attempts to exceed the maximum limitation, the current in the L.P. speed limiter and temperature control solenoid is reduced. This opens the spill valve to reduce the pressure on the pressure drop control diaphragm. The flow control spill valve then opens to reduce the pump servo pressure and fuel pump output.

To prevent the L.P. compressor from overspeeding, multi-spool engines usually have an L.P. compressor shaft speed governor. A signal of L.P. shaft speed and intake temperature is fed to an amplifier and solenoid valve, the valve limiting the fuel flow in the same way as the gas temperature control.

The system described uses main and starting spray nozzles under the control of an H.P. shutoff valve. Two starting nozzles are fitted in the combustion chamber, each being forward of an igniter plug. When the engine has started, the fuel flow to these nozzles is cut off by the H.P. shutoff valve.

To ensure that a satisfactory fuel pressure to the spray nozzles is maintained at high altitudes, a back pressure valve, located downstream of the throttle valve, raises the pressure levels sufficiently to ensure satisfactory operation of the fuel pump servo system.

Flow Control

A flow control fuel system is generally more compact than a pressure control system and is not sensitive to flow effect of variations downstream of the throttle. The fuel pump delivery pressure is related to engine speed; thus, at low engine speeds pump delivery pressure is quite low. The fuel pump output is controlled to give a constant pressure difference across the throttle valve at a constant air intake condition. Various devices are also used to adjust the fuel flow for air intake pressure variations, idling and acceleration control, gas temperature and compressor delivery pressure control.

A variation of the flow control system is the proportional flow control system (Figure 7–7), which is more suitable for engines requiring large fuel flows and which also enables the fuel trimming devices to adjust the fuel flow more accurately. A small controlling flow is created that has the same characteristics as the main flow, and this controlling or proportional flow is used to adjust the main flow.

A different type of spill valve, referred to as a kinetic valve, is used in this system. This valve consists of two opposing jets, one subjected to pump delivery pressure and the other to pump servo pressure, and an interrupter blade that can be moved between the jets (Figure 7–8). When the blade is clear of the jets, the kinetic force of the H.P. fuel jet causes the servo pressure to rise (spill valve closed) and the fuel pump moves to maximum stroke to increase the fuel flow. When the blade is lowered between the jets, the pressure jet is deflected and the servo pressure falls, so reducing the pump stroke and the fuel flow. When the engine is steadily running, the blade is in an intermediate position allowing a slow bleed from servo and thus balancing the fuel pump output.

All the controlling devices, except for the engine speed governor, are contained in one combined fuel control unit. The main parts of the control unit are the altitude sensing unit (A.S.U.), the acceleration control unit (A.C.U.), the throttle and pressurizing valve unit, and the proportioning valve unit.

At any steady running condition below governed speed, the fuel pump delivery is controlled to a fixed value by the A.S.U. The spill valve in this unit is held in the controlling position by a balance of forces, spring force and the piston force. The piston is sensitive to the pressure difference across the sensing valve, the pressure difference being created by fuel flowing from the proportioning valve back to the fuel pump inlet.

The proportioning valve diaphragm is held open in a balanced condition allowing fuel to pass to the A.S.U. This means that the restrictor outlet pressure is equal to the throttle outlet pressure and, as their inlet pressures are equal, it follows that the pressure difference across the restrictors and the throttle are equal; therefore, a constant fuel flow is obtained.

When the throttle is slowly opened, the pressure difference across the throttle valve and the proportioning flow restrictors decreases and the proportioning valve diaphragm adjusts its position. This reduces the proportional flow, which closes the A.S.U. spill valve and increases the servo pressure. The fuel pump increases its delivery and this restores the pressure difference across the throttle valve and equalizes the pressure difference across the restrictors. The proportional flow is restored to its original value and the balance of forces in the A.S.U. returns the spill valve to the controlling position.

A variation of air intake pressure, due to a change of aircraft forward speed or altitude, is sensed by the capsule in the A.S.U. A pressure reduction causes the A.S.U. capsule to expand, thus increasing the bleed from the spill valve. This reduces fuel pump delivery until the fuel flow matches the airflow and results in a lower pressure difference across the throttle valve and the proportioning valve restrictors. The reduced proportional flow restores the balance in the A.S.U. that returns the spill valve to its controlling position. Conversely, an increase in aircraft forward speed increases the air intake pressure, which reduces the bleed from the spill valve and increases the fuel flow.

During a rapid acceleration, the sudden decrease in throttle pressure difference is sensed by the A.S.U., causing the spill valve to close. Such a rapid increase in fuel supply would, however, create an excessive gas temperature and also cause the compressor to surge. This occurs because the inertia of the rotating assembly results in an appreciable time lag in the rate of airflow increase. It is essential, therefore, to have an acceleration control to override the A.S.U. to give a corresponding lag in the rate of fuel flow increase.

The rapid initial increase of fuel flow causes a rise in the pressure difference across the fuel metering plunger and this is sensed by a diaphragm in the pressure drop control section. At a fixed value of overfueling, the pressure drop control diaphragm opens its servo spill valve to override the A.S.U. and maintains a constant pressure difference across the metering plunger.

The increased fuel supply causes the engine to accelerate and the fuel metering plunger gives the maximum permissible fuel flow to match the increasing compressor delivery pressure. This it achieves through the A.C.U. servo system, which is under the control of a spill valve operated by compressor delivery air pressure acting on a capsule.

As the compressor delivery pressure continues to rise, the capsule is compressed to open the spill valve and to bleed pressure from above the metering plunger. Pump delivery pressure acting underneath the plunger causes it to lift, this increases the area of the main fuel flow passage.

The pressure drop control spill valve closes to increase the fuel pump delivery and maintains the controlling pressure difference across the plunger. The fuel flow, therefore, progressively rises as airflow through the compressor increases. The degree of overfueling can be automatically changed by the air switch, which increases the pressure signal on to the capsule. The full value of compressor delivery pressure is now passed on to the A.C.U. capsule assembly, thus increasing the opening rate of the metering plunger.

As the controlled overfueling continues, the pressure difference across the throttle valve increases. When it reaches the controlling value, the A.S.U. takes over due to the increasing proportional flow and again gives a steady fuel flow to the spray nozzles.

The engine speed governor can be of the pressure control type described previously, or a hydro-mechanical governor also described previously.

The control of servo pressure by the hydro-mechanical governor is very similar to that of the pressure control governor, except that the governor pressure is obtained from pump delivery fuel passing through a restrictor and the restricted pressure is controlled by a rotating spill valve; this type of governor is unaffected by changes in fuel specific gravity.

At low engine speeds, the rotating spill valve is held open; however, as engine speed increases, centrifugal loading moves the valve towards the closed position against the diaphragm loads. This restricts the bleed of H.P. fuel to the L.P. side of the drum until, at governed speed, the governor pressure deflects the diaphragm and opens the fuel pump servo pressure spill valve to control the maximum fuel flow and engine speed.

If the engine gas temperature exceeds its maximum limitation, the solenoid on the proportioning valve unit is progressively energized. This causes a movement of the rocker arm to increase the effective flow area of one restrictor, thus increasing the proportional flow and opening the A.S.U. spill valve to reduce servo pressure. The fuel flow is thus reduced and any further increase of gas temperature is prevented.

To prevent the L.P. compressor from overspeeding, some twin-spool engines have an L.P. shaft rpm governor. A signal of L.P. shaft speed is fed to an amplifier and solenoid valve, which limits the fuel output in the same way as the gas temperature control.

An idling speed governor is often fitted to ensure that the idling rpm does not vary with changing engine loads. A variation of idling rpm causes the rocker arm to move and alter the proportional flow, and the A.S.U. adjusts the pump delivery until the correct idling rpm is restored.

On some engines, a power limiter is used to prevent overstressing of the engine. To achieve this, compressor delivery pressure acts on the power limiter capsule. Excess pressure opens the power limiter atmospheric bleed to limit the pressure on the A.C.U. capsule and this controls the fuel flow through the metering plunger.

To enable the engine to be relit and to prevent flame-out at altitude, the engine idling rpm is made to increase with altitude. To achieve this, some engines incorporate a minimum flow valve that adds a constant minimum fuel flow to that passing through the throttle valve.

Combined Acceleration and Speed Control

The combined acceleration and speed control system (Figure 7–9) is a mechanical system without small restrictors or spill valves. It is also an all-speed governor system and therefore needs no separate governor unit for controlling the maximum rpm. The controlling mechanism is contained in one unit, usually referred to as the fuel flow regulator (F.F.R.). An H.P. fuel pump is used and the fuel pump servo piston is operated by H.P. fuel on one side and main spray nozzle (servo) pressure on the spring side.

The F.F.R. is driven by the engine through a gear train and has two centrifugal governors, known as the speed control governor and the pressure drop control governor. Two sliding valves are also rotated by the gear train. One valve, known as the variable metering sleeve, has a triangular orifice, known as the variable metering orifice (V.M.O.), and this sleeve is given axial movement by a capsule assembly. The V.M.O. sleeve moves inside a non-rotating governor sleeve that is moved axially by the speed control governor. The other valve, known as the pressure drop control valve, is provided with axial movement by the pressure drop control governor and has a triangular orifice, known as the pressure drop control orifice, and a fixed-area rectangular orifice. The speed control governor is set by the throttle lever through a cam, a spring, and a stirrup arm inside the regulator.

At any steady running condition, the engine speed is governed by the regulator controlling the fuel flow. The fuel pump delivery is fixed at a constant value by applying the system pressure difference to the fuel pump servo piston. This is arranged to balance the servo piston spring forces.

When the air intake pressure is at a constant value, the rotating V.M.O. sleeve is held in a fixed axial position by the capsule loading. The fixed throttle setting maintains a set load on the speed control governor and, as the rpm is constant, the governor sleeve is held in a fixed position.

The fuel pump delivery is passed to the annulus surrounding the V.M.O.; the annulus area is controlled by the governor sleeve, and the exposed area of the orifice is set by the axial position of the V.M.O. sleeve. Consequently, fuel passes to the inside of the sleeve at a constant flow and therefore at a constant pressure difference.

The pressure drop control valve, which also forms a piston, senses the pressure difference across the V.M.O. and maintains the fuel flow at a fixed value in relation to a function of engine speed, by controlling the exposed area of the pressure drop control orifice.

When the throttle is slowly opened, the load on the speed control governor is increased, so moving the governor sleeve to increase the V.M.O. annulus area. The effect of opening the V.M.O. is to reduce the pressure difference and this is sensed by the pressure drop control governor, which opens the pressure drop valve. The reduced system pressure difference is immediately sensed by the fuel pump servo piston, which increases the pump stroke and consequently the fuel output. The increased compressor delivery pressure acts on the capsule assembly that gradually opens the V.M.O. so that the fuel flow and engine speed continue to increase. At the speed selected, centrifugal forces acting on the speed control governor move the governor sleeve to reduce the V.M.O. annulus area. The resultant increased pressure difference is sensed by the pressure drop control governor, which adjusts the pressure drop valve to a point at which the pump servo system gives an output to match the engine requirements. The function of the governors and the control of the fuel flow is shown diagrammatically in Figure 7–10.

During a rapid acceleration, the initial degree of overfueling is mechanically controlled by a stop that limits the opening movement of the speed control governor sleeve. A similar stop also prevents the fuel supply from being completely cut off by the governor sleeve during a rapid deceleration.

Changes in altitude or forward speed of the aircraft vary the fuel flow required to maintain a constant engine speed. To provide this control, the capsule assembly senses changes in H.P. compressor inlet and delivery pressures and adjusts the V.M.O. accordingly. For instance, as the aircraft altitude increases, the compressor delivery pressure falls and the capsule assembly expands to reduce the V.M.O. The increased system pressure drop is sensed by the fuel pump servo piston, which adjusts the pump output to match the reduced airflow and so maintain a constant engine speed. Conversely, an increase in aircraft forward speed causes the capsule assembly to be compressed and increase the V.M.O. The reduced system pressure drop causes the fuel pump to increase its output to match the increased airflow.

To prevent the maximum gas temperature from being exceeded, fuel flow is reduced in response to signals from thermocouples sensing the temperature. When the maximum temperature is reached, the signals are amplified and passed to a rotary actuator, which adjusts the throttle mechanism. This movement has the same effect on fuel flow as manual operation of the throttle.

To ensure that the engine is not overstressed, the H.P. compressor delivery pressure is controlled to a predetermined value. At this value, a pressure-limiting device, known as a power limiter, reduces the pressure in the capsule chamber, thus allowing the capsule assembly to expand and reduce the V.M.O. so preventing any further increase in fuel flow.

A governor prevents the L.P. compressor shaft from exceeding its operating limitations and also acts as a maximum speed governor in an event of a failure of the F.F.R. The governor provides a variable restrictor between the regulator and the main fuel spray nozzle manifold. Should the L.P. compressor reach its speed limitation, flyweights in the governor move a sleeve valve to reduce the flow area. The increased system pressure drop is sensed by the fuel pump servo piston, which reduces the fuel flow to the spray nozzles.

This fuel system has no pressurizing valve to divide the flow from the fuel pump into main and primary fuel flows. Primary fuel pressure is taken from the fixed-area orifice of the pressure drop control valve. This pressure is always higher than the main fuel pressure and it is not shut off by the pressure drop control piston. It therefore gives a satisfactory idling fuel flow at all altitudes.

On engines featuring water injection, a reset device (Figure 7–11), operated by a piston and reset cam, increases the loading on the throttle control spring and stirrup arm, thus selecting a higher engine speed during water injection. To prevent the power limiter (Figure 7–9) from canceling the effect of water injection, a capsule in the limiter is subjected to water pressure to raise the compressor delivery pressure at which the power limiter operates.

Pressure Ratio Control

The pressure ratio control (Figure 7–12) is a mechanical system similar to the combined acceleration and speed control system, but uses the ratio of H.R. compressor delivery pressure to air intake pressure (1¾/P₁) as the main controlling parameter. It needs no separate governor unit for controlling the maximum rpm The controlling mechanism is contained in one unit, which is usually referred to as a fuel flow regulator (F.F.R.). A gear-type pump is used, as described later, and the pump output to the F.F.R. is controlled by a pressure drop spill valve.

The F.F.R. is driven by the engine through a gear train and has two rotating valves. One valve, known as a variable metering sleeve, has a triangular orifice, known as the variable metering orifice (V.M.O.), and this sleeve is given axial movement by a capsule assembly. The other valve, known as the pressure drop control valve, is provided with axial movement by a centrifugal governor, known as a pressure drop control governor. Both valves form variable restrictors which control the fuel flow to the spray nozzles.

Control of the V.M.O. area is a function of a pressure ratio control unit housed in the F.F.R. A pressure ratio control valve, subjected to P₄ and P₁ pressures, regulates the movement of the F.F.R. capsule and thus controls the V.M.O. area to produce the pressure ratio dictated by the throttle or power lever.

At any steady running condition, the output of the fuel pump is greater than the engine requirement. The pressure drop spill valve is open to allow surplus fuel to return to the inlet side of the pump. This action controls the fuel delivery to that demanded by the F.F.R.

When the throttle is slowly opened, the throttle-controlled orifice is increased and the control pressure falls, thus allowing the pressure ratio control valve to move towards the closed position (acceleration stop). F.F.R. capsule chamber pressure increases and the capsule moves the metering sleeve to increase the V.M.O. area. The effect of opening the V.M.O. is to reduce the pressure difference and this is sensed by the pressure drop governor, which opens the pressure drop control orifice. The reduced system pressure difference is immediately sensed by the pressure drop spill valve, which moves towards the closed position and consequently increases the fuel output. The increased fuel flow accelerates the engine with a subsequent increase in pressure ratio (P₄/P₁). When the required pressure ratio is reached, the pressure ratio control valve opens and the F.F.R. capsule chamber pressure reduces. The capsule assembly expands, moving the V.M.O. sleeve to reduce the orifice area. The resultant increased pressure difference is sensed by the pressure drop control governor, which adjusts the pressure drop control orifice to a point at which the pressure drop spill valve gives a fuel output consistent with steady running requirements.

During a rapid acceleration, the degree of overfueling is mechanically controlled by the acceleration stop, which limits the movement of the pressure ratio control valve. A similar stop prevents the fuel supply from being completely cut off during a rapid deceleration.

When accelerating to a higher P₄/P₁ ratio, the throttle control orifice is increased. The reduced pressure allows the pressure ratio control capsule to contract so that the valve contacts the acceleration stop. F.F.R. capsule chamber pressure increases and the capsule moves to increase the V.M.O. area. This action continues until the required P₄/P₁ ratio is reached. The increased P₄ pressure allows the pressure ratio control capsule to re-expand and the valve to return to the steady running position.

A change in altitude of the aircraft requires a variation in fuel flow to match the engine thrust and aircraft climb requirement. The normal effect of an altitude increase is to decrease the P₁ and P₄ pressures, thus opening the pressure ratio control valve and allowing the F.F.R. capsule to expand to reduce the V.M.O. area and, in consequence, the fuel flow. However, to match the engine thrust and aircraft climb requirement it is necessary to increase the P₄/P₁ ratio with increasing altitude. This is done by a trimmer valve and a capsule that is subjected to P₁ pressure. As P₁ pressure decreases, the trimmer valve moves across the P₁ controlled orifice to reduce the control pressure. This is sensed by the control capsule, which, by acting on the pressure ratio control valve, slows the closure of the V.M.O. as altitude is increased. This maintains the thrust requirement with the throttle at a fixed position.

To prevent the maximum L.P. compressor rpm and engine gas temperature from being exceeded, a valve known as the auxiliary throttling valve is fitted in the outlet from the fuel pump. Under steady running conditions, the valve is held open by spring force. When limiting conditions are reached, the fuel flow is reduced in response to speed and temperature signals from the engine. The signals are amplified and passed to a rotary actuator that reduces the area of a variable restrictor. The effect of this is to increase the fuel pressure, which partially closes the throttling valve. H.P. fuel pressure acting on the face of the pressure drop spill valve is increased and the spill valve opens to reduce the fuel flow to the spray nozzles.

H.P. shaft speed is also governed by the auxiliary throttling valve. Should other controlling devices fail and pump speed increase, the fuel pressure closes the throttling valve and opens the pressure drop spill valve to reduce the fuel flow.

With the throttle closed, idling condition is determined by controlling the amount of air being vented through the idling adjuster and the ground idling solenoid valve. With both bleeds in operation, satisfactory flight idling for the air off-takes is ensured. By closing the solenoid valve a lower power condition for ground idling is obtained.

This fuel system, like the combined acceleration and speed control system, has no pressurizing valve to divide the flow from the fuel pump into main and primary flows.

Speed and Temperature Control Amplifiers

The speed and temperature control amplifier receives signals from thermocouples measuring E.G.T. and from speed probes sensing L.P. and in some cases I.P. shaft speeds (N₁ and N₂). The amplifier basically comprises speed and temperature channels that monitor the signals sensed. If either N₁, N₂, or E.G.T. exceeds pre-set datums, the amplifier output stage is triggered to connect an electrical supply to a solenoid valve or a variable restrictor that overrides the F.F.R. and causes a reduction in fuel flow. The limiter will only relinquish control back to the F.F.R. if the input conditions are altered (altitude, speed, ambient temperature, or throttle lever position). The limiter system is designed to protect against parameters exceeding their design values under normal operation and basic fuel system failures.

Engine Supervisory Control

The engine supervisory control (E.S.C.) system performs a supervisory function by trimming the fuel flow scheduled by the fuel flow governor (F.F.G.) to match the actual engine power with a calculated engine power for a given throttle angle. The E.S.C. provides supervisory and limiting functions by means of a single control output signal to a torque motor in the F.F.G. In order to perform its supervisory function the E.S.C. monitors inputs of throttle angle, engine bleed state, engine pressure ratio (E.P.R.), and air data computer information (altitude, Mach number, and temperatures). From this data the supervisory channel predicts the value of N₁ required to achieve the command E.P.R. calculated for the throttle angle set by the pilot. Simultaneously a comparison is made between the command E.P.R. and the actual E.P.R. and the difference is compared with a programmed datum.

During acceleration the comparitor connects the predicted value of N₁ to the limiter channel until the difference between the command and actual E.P.R. is approximately 0.03 E.P.R. At this point the predicted L.P. shaft speed is disconnected and the E.P.R. difference signal is connected to the limiter channel.

The final output from the supervisory channel, in the form of an error signal, is supplied to a “lowest wins” circuit along with the error signals from the limiter channel. While the three error signals remain positive (N₁ and E.G.T. below datum level and actual E.P.R. below command E.P.R.) no output is signaled to the torque motor. If, however, the output stage of the E.S.C. predicts that E.G.T. will exceed datum or that N₁ will either exceed its datum or the predicted level for the command E.P.R., then a signal is passed to the torque motor to trim the fuel flow.

Plunger-Type Fuel Pump

The pump shown in Figure 7–14 is of the single-unit, variable-stroke, plunger type; similar pumps may be used as double units depending upon the engine fuel flow requirements.

The fuel pump is driven by the engine gear train and its output depends upon its rotational speed and the stroke of the plungers. A single-unit fuel pump can deliver fuel at the rate of 100–2000 gallons per hour at a maximum pressure of about 2000 lb. per square inch. To drive this pump, as much as 60 horsepower may be required.

The fuel pump consists of a rotor assembly fitted with several plungers, the ends of which project from their bores and bear on to a non-rotating camplate. Due to the inclination of the camplate, movement of the rotor imparts a reciprocating motion to the plungers, thus producing a pumping action. The stroke of the plungers is determined by the angle of inclination of the camplate. The degree of inclination is varied by the movement of a servo piston that is mechanically linked to the camplate and is biased by springs to give the full stroke position of the plungers. The piston is subjected to servo pressure on the spring side and on the other side to pump delivery pressure; thus variations in the pressure difference across the servo piston cause it to move with corresponding variations of the camplate angle and, therefore, pump stroke.

Gear-Type Fuel Pump

The gear-type fuel pump (Figure 7–12) is driven from the engine and its output is directly proportional to its speed. The fuel flow to the spray nozzles is controlled by recirculating excess fuel delivery back to inlet. A spill valve, sensitive to the pressure drop across the controlling units in the system, opens and closes as necessary to increase or decrease the spill.

Fuel Spray Nozzles

The final components of the fuel system are the fuel spray nozzles, which have as their essential function the task of atomizing or vaporizing the fuel to ensure its rapid burning. The difficulties involved in this process can be readily appreciated when one considers the velocity of the air stream from the compressor and the short length of combustion system in which the burning must be completed.

An early method of atomizing the fuel is to pass it through a swirl chamber where tangentially disposed holes or slots imparted swirl to the fuel by converting its pressure energy to kinetic energy. In this state, the fuel is passed through the discharge orifice that removes the swirl motion as the fuel is atomized to form a cone-shaped spray. This is called “pressure jet atomization.” The rate of swirl and pressure of the fuel at the fuel spray nozzle are important factors in good atomization. The shape of the spray is an indication of the degree of atomization as shown in Figure 7–15. Later fuel spray nozzles utilize the airspray principle that employs high velocity air instead of high velocity fuel to cause atomization. This method allows atomization at low fuel flow rates (provided sufficient air velocity exists) thus providing an advantage over the pressure jet atomizer by allowing fuel pumps of a lighter construction to be used.

The atomizing spray nozzle, as distinct from the vaporizing burner, has been developed in five fairly distinct types, the Simplex, the variable port (Lubbock), the Duplex or Duple, the spill type, and the airspray nozzle.

The Simplex spray nozzle shown in Figure 7–16 was first used on early jet engines. It consists of a chamber, which induces a swirl into the fuel, and a fixed-area atomizing orifice. This fuel spray nozzle gave good atomization at the higher fuel flows, that is, at the higher fuel pressures, but was very unsatisfactory at the low pressures required at low engine speeds and especially at high altitudes. The reason for this is that the Simplex was, by the nature of its design, a “square law” spray nozzle; that is, the flow through the nozzle is proportional to the square root of the pressure drop across it. This meant that if the minimum pressure for effective atomization was 30 lb. per square inch, the pressure needed to give maximum flow would be about 3000 lb. per square inch. The fuel pumps available at that time were unable to cope with such high pressures so the variable port spray nozzle was developed in an effort to overcome the square law effect.

Although now only of historical value, the variable port or Lubbock fuel spray nozzle (Figure 7–17) made use of a spring-loaded piston to control the area of the inlet ports to the swirl chamber. At low fuel flows, the ports were partly uncovered by the movement of the piston; at high flows, they were fully open. By this method, the square law pressure relationship was mainly overcome and good atomization was maintained over a wide range of fuel flows. The matching of sets of spray nozzles and the sticking of the sliding piston due to dirt particles were, however, difficulties inherent in the design, and this type was eventually superseded by the Duplex and the Duple fuel spray nozzles.

The Duplex and the Duple spray nozzles require a primary and a main fuel manifold and have two independent orifices, one much smaller than the other. The smaller orifice handles the lower flows and the larger orifice deals with the higher flows as the fuel pressure increases. A pressurizing valve may be employed with this type of spray nozzle to apportion the fuel to the manifolds (Figure 7–18). As the fuel flow and pressure increase, the pressurizing valve moves to progressively admit fuel to the main manifold and the main orifices. This gives a combined flow down both manifolds. In this way, the Duplex and Duple nozzles are able to give effective atomization over a wider flow range than the Simplex spray nozzle for the same maximum fuel pressure. Also, efficient atomization is obtained at the low flows that may be required at high altitude. In the combined acceleration and speed control system, the fuel flow to the spray nozzles is apportioned in the F.F.R.

The spill-type fuel spray nozzle can be described as being a Simplex spray nozzle with a passage from the swirl chamber for spilling fuel away. With this arrangement it is possible to supply fuel to the swirl chamber at a high pressure all the time. As the fuel demand decreases with altitude or reduction in engine speed, more fuel is spilled away from the swirl chamber, leaving less to pass through the atomizing orifice. The spill spray nozzles’ constant use of a relatively high pressure means that even at the extremely low fuel flows that occur at high altitude there is adequate swirl to provide constant and efficient atomization of the fuel.

The spill spray nozzle system, however, involves a somewhat modified type of fuel supply and control system from that used with the previous types. A means has to be provided for removing the spill and also for controlling the amount of spill flow at various engine operating conditions. A disadvantage of this system is that excess heat may be generated when a large volume of fuel is being recirculated to inlet. Such heat may eventually lead to a deterioration of the fuel.

The airspray nozzle (Figure 7–19) carries a proportion of the primary combustion air with the injected fuel. By aerating the spray, the local fuel-rich concentrations produced by other types of spray nozzle are avoided, thus giving a reduction in both carbon formation and exhaust smoke. An additional advantage of the airspray nozzle is that the low pressures required for atomization of the fuel permit the use of the comparatively lighter gear-type pump.

A flow distributor (Figure 7–20) is often required to compensate for the gravity head across the manifold at low fuel pressures to ensure that all spray nozzles pass equal quantities of fuel.

Some combustion systems vaporize the fuel as it enters the combustion zone.

Fuel Requirements

In general, a gas turbine fuel should have the following qualities:

The pumping qualities of the fuel depend upon its viscosity or thickness, which is related to fuel temperature. Fuel must be satisfactory down to approximately –50°C. As the fuel temperature falls, ice crystals may form to cause blockage of the fuel filter or the orifices in the fuel system. Fuel heating and anti-icing additives are available to alleviate this problem.

For easy starting, the gas turbine engine depends upon the satisfactory ignition of the atomized spray of fuel from the fuel spray nozzles, assuming that the engine is being motored at the required speed. Satisfactory ignition depends upon the quality of fuel in two ways:

The calorific value (Figure 7–21) of a fuel is an expression of the heat or energy content per pound or gallon that is released during combustion. This value, which is usually expressed in British thermal units, influences the range of an aircraft. Where the limiting factor is the capacity of the aircraft tanks, the calorific value per unit volume should be as high as possible, thus enabling more energy, and hence more aircraft range, to be obtained from a given volume of fuel. When the useful payload is the limiting factor, the calorific value per unit of weight should be as high as possible, because more energy can then be obtained from a minimum weight of fuel. Other factors that affect the choice of heat per unit of volume or weight must also be taken into consideration; these include the type of aircraft, the duration of flight, and the required balance between fuel weight and payload.

Turbine fuels tend to corrode the components of the fuel and combustion systems mainly as a result of the sulfur and water content of the fuel. Sulfur, when burned in air, forms sulfur dioxide; when mixed with water this forms sulfurous acid and is very corrosive, particularly on copper and lead. Because it is impracticable to completely remove the sulfur content, it is essential that the sulfur be kept to a controlled minimum. Although free water is removed prior to use, dissolved water, i.e., water in solution, cannot be effectively removed, as a fuel would reabsorb moisture from the atmosphere when stored in a vented aircraft or storage tank.

All gas turbine fuels are potentially dangerous; handling storage precautions should be strictly observed.

Vapor Locking and Boiling

The main physical difference between kerosene and wide-cut fuels is their degree of volatility, the latter type of fuel having a higher volatility, thus increasing the problem of vapor locking boiling. With kerosene-type fuels, the volatility is controlled by distillation and flash point, but with the wide-cut fuels it is controlled by distillation Reid Vapor Pressure (R.V.P.) test. In this test, the absolute pressure of the fuel is recorded by special apparatus with the fuel temperature at 37.8°C (100°F).

Kerosene has a low vapor pressure and will boil only at extremely high altitudes or high temperatures, whereas a wide-cut fuel will boil at a much lower altitude.

The fuel temperature during flight depends upon altitude, rate of climb, duration at altitude, and kinetic heating due to forward speed. When boiling does occur, the vapor loss can be very high, especially with wide-cut fuels, and this may cause vapor locking with consequent malfunctions of the engine fuel system and fuel metering equipment.

To obviate or reduce the risk of boiling, it is usual to pressurize the fuel tanks. This involves maintaining an absolute pressure above the fuel in excess of its vapor pressure at any specific temperature. This may be accomplished by using an inert gas or by using the fuel vapor pressure with a controlled venting system.

For sustained supersonic flight, some measure of tank insulation is necessary to reduce kinetic heating effects, even when lower volatility fuels are used.

Fuel Contamination Control

Fuel can be maintained in good condition by well-planned storage and by making routine aircraft tank drain checks. The use of suitable filters, fuel/water separators, and selected additives will restrict the contamination level, e.g., free water and solid matter, to a practical minimum. Keeping the fuel free of undissolved water will prevent serious icing problems, reduce the microbiological growth, and minimize corrosion. Reducing the solid matter will prevent undue wear in the fuel pumps, reduce corrosion, and lessen the possibility of blockage occurring within the fuel system.

Combustion Process

The combustion process primarily entails the exothermic reaction of a hydrocarbon fuel with oxygen, to produce carbon dioxide and water. The number of moles of each is dictated by the fuel composition. For illustration the reaction for one of the hydrocarbons present in kerosene is described below.

The stoichiometric fuel air ratio, where all the atmospheric oxygen would he consumed, is then easily derived. The required number of moles of air is evaluated, and then converted to mass. This is shown below using the molar oxygen content of air, and then the molecular weight of air:

15 × (1/0.2095) = 71.6 Number of moles of air

71.6 × 28.964 = 2073.8 Mass of air

140/2073.8 = 0.0675 FAR by mass

In fact the chemical reaction is far more complex with many reactions taking place simultaneously. This includes the above for the whole range of hydrocarbons present, as well as formation and consumption of carbon monoxide and oxides of nitrogen (NO_x). The last is due to dissociation of atmospheric nitrogen.

Direct Firing and Indirect Firing

The vast majority of gas turbine engines employ direct firing where the fuel is injected into the engine combustion chamber and then burned. This is the case for the three primary fuels: kerosene, diesel, and natural gas, and also for a number of less common fuels.

However, for certain obscure fuels considered for industrial engines direct firing may not be practical due to their corrosive or erosive nature. The only manner in which such fuel may be utilized is via indirect firing where it is burned external to the gas turbine, with a heat exchanger to transfer heat to the compressor delivery air. Another indirect firing case of interest is a nuclear powered closed cycle.

For indirect firing the following changes must be made to performance calculations:

The impact of both of the above is to lower power output and thermal efficiency since the mass flow and CP in the turbines are both lower. The exact power loss depends upon the engine cycle, and is between 4 and 8%.

Kerosene

Kerosene is a faction of crude oil primarily comprising a band of hydrocarbons with an average composition of C₁₂H_23.5 and molecular weight of 167.7. There are a number of commercial grades available such as JP4, JP5, Jet Al, and AVTUR that are refined to a tight specification. The proprietary high cost JP10 is a high-density fuel used for certain military applications such as missiles. The term kerosene is usually also taken to encompass the tighter faction aviation gasoline, or AVGAS.

Aeroengines almost exclusively use kerosene as its high calorific value minimizes fuel weight, and because it is free of corrosive elements such as sulfur, which is essential for such high cost engines. The premium in fuel price is warranted by these gains.

Diesel

Diesel fuel is a heavier faction of crude oil than kerosene, again comprising a band of hydrocarbons, with an average composition of C_12.9H_23.9 and molecular weight of 178.6. It is less refined and hence has a lower cost, but therefore also contains small percentages of other elements such as the corrosive agent sulfur. The terms diesel and fuel oil are commonly used interchangeably. There are a number of grades spanning Number 1 fuel oil to Number 6 fuel oil, the tolerance relative to the nominal specification for each grade is significantly wider than for the kerosenes. The lower Numbers are closer to kerosene, have lower amounts of sulfur (typically up to 1.5% by weight being acceptable), and hence usually only Numbers 1 and 2 are used for gas turbine engines.

For cost considerations, diesel is almost exclusively used for marine engines, and in military applications it presents a lower risk of explosion. Its corrosive sulfur content is less of an issue, since marine engines must be designed to withstand a corrosive atmosphere anyway due to the seawater environment.

Though land-based power generation engines usually burn natural gas, a back-up liquid fuel capability is often required to guard against an interrupted gas supply. In addition some niche applications use only liquid fuel. Again, due to cost, fuel oil Numbers 1 or 2 are employed, but life is shorter than for natural gas.

Natural Gas

Natural gas comprises over 80% methane with minor amounts of ethane, propane, butane, and heavier hydrocarbons. It may also include carbon dioxide, nitrogen, and hydrogen. There are a plethora of blends of natural gas available worldwide. The composition on a molar basis of a common blend, used as a “typical” natural gas for deriving gas properties of combustion products, is listed below:

As stated, industrial engines for power generation predominantly use natural gas. This is because of its abundance, competitive cost, and negligible content of corrosive elements such as sulfur. The benefits include lower emissions of carbon dioxide and good engine life, and the large volume required is not a logistics issue as it is for aircraft or marine propulsion. The lower emissions of carbon dioxide are because the lighter hydrocarbons contain more hydrogen and produce more water and less carbon dioxide when burned. As described later the higher water content of the combustion products leads to a higher power output and thermal efficiency for a given engine and SOT level. Engines for natural gas pumping burn solely the gas tapped off the pipeline.

Liquefied petroleum gas (LPG) may be burned in gas turbines, injected either as a liquid or as a gas. Refinery tail gases are also of interest, and are produced as a by-product of oil refining. Other gas fuels occur naturally in lower quantities, such as condensates and naphtha. Their composition does not comprise the high percentage of methane present in “natural gas” and their properties provide significant difficulties for satisfactory fuel injection. For instance, condensates condense into a liquid at ambient temperature under the action of pressure.

Other Fuels

For industrial engines other fuels have been considered for specific applications/projects, but none have managed to take a significant share of the market.

Extensive efforts over many decades have been spent on developing coal as a fuel for gas turbine engines; however, to date success has been limited. Coal gasification involves pyrolysis of coal in an atmosphere of steam and usually oxygen (as opposed to air). The resultant fuel comprises primarily carbon monoxide and hydrogen, with minor amounts of carbon dioxide, methane, and sulfur compounds. Other attempts have also been made to coarse mill the coal to around 1 mm particles and then chemically clean it to remove the corrosive and erosive elements such as sodium, potassium, vanadium, and ash (silicon-based compounds). It is fine milled to sub-10 μm particle size such that it has characteristics of a gas for direct firing.

Hydrogen has been considered on a research basis for aircraft engines due to its high calorific value by weight. However, the volume of fuel required is prohibitive, requiring complex tank arrangements. Hydrogen is present in many blends of natural gas; too high a content leads to problems with fuel injector flow number.

Biomass is a gas fuel produced from natural vegetation. While it has been considered more recently as a renewable fuel it is not widely used due to the large area of cultivated land required to fuel one engine continuously.

Calorific or Heating Value on a Mass or Volumetric Basis

Calorific value is the heat energy released per unit mass of fuel burned as shown by Formula F7.1. Gross calorific value, or higher heating value (HHV), is the total heat released making no allowance for the latent heat required to vaporize the liquid water produced by the combustion process. Hence it is a theoretical parameter and is rarely used. The net calorific value, or lower heating value (LHV), does account for the latent heat of vaporization and hence is the parameter most commonly used for gas turbine performance calculations. Strictly LHV is quoted as the heat released under pressure in a constant volume when the combustion products are cooled to the initial temperature of 25°C. Hence for exhaustively rigorous combustion calculations the enthalpy balance should include:

Occasionally for gas fuels calorific value is quoted on a volumetric basis (Formula F7.2) having units of either kJ/m³ or kl/scm. The scm (standard cubic meter) is the amount of gas that would occupy 1 m³ at the ISO pressure and temperature of 101.325 kPa and 288.15 K respectively. Formula F7.3 shows how the number of scm present in a volume of gas fuel at any given pressure and temperature may be calculated.

LHV is used in most performance calculations such as design point, off design and test data analysis. Nominal values for the three primary gas turbine fuels are presented below, followed by methodologies for deriving values for a specific batch of fuel used during an engine test.

Density and Specific Gravity

Fuel density, the mass of fuel in a given volume (Formula F7.4), is important during engine performance testing. Invariably volumetric flow rate is measured and density must be derived such that heat release may be calculated. It is not of interest for other gas turbine performance topics.

Specific gravity is commonly used for liquid fuels and is the ratio of the fuel density to that of water at 4°C as shown by Formula F7.5. Figure 7–22 presents SG versus fuel temperature for commonly used kerosenes and diesels. At 15°C the SG for AVGAS is 0.704, the rest of the kerosenes are between 0.76 and 0.818, and diesels range from 0.82–0.88. For a given kerosene type the tolerance is around ±0.1 and for diesel it is even wider. Hence Figure 7–22 can only be used as indicative of SG level and for key engine performance tests it must be measured. This is accomplished using a fuel sample in a laboratory with a calibrated hydrometer. The fuel sample temperature must be measured simultaneously and Formula F7.6 is used to derive SG at the fuel temperature measured during the engine performance test.

Gas fuel density may be calculated from Formula F7.7 where the “compressibility” term z is used to modify the perfect gas law such that it may be applied to non-perfect gases. Gas fuel pressure and temperature must be measured in the vicinity of the turbine flow meter. ASME (1993) Performance Test Code PTC–22 and ASME (1969) Gaseous Fuels PTC 3.3 show how R and z may be derived once the fuel gas composition is known. The calculation process for the latter is complex; however, if it is not followed then significant errors will result.

Viscosity, Dynamic and Kinematic

Dynamic viscosity is the resistance to movement of one layer of a fluid over another and is defined by Formula F7.8. Kinematic viscosity is dynamic viscosity divided by density (Formula F7.9) and is the ratio of viscous forces to inertia forces. Dynamic viscosity for liquid fuels is occasionally required for performance calculations in determining fuel pump power requirements such as during start modeling. Kinematic viscosity is a second-order variable in the calibration of bulk meters and turbine flow meters for measuring liquid fuel volumetric flow rate.

Figure 7–23 presents kinematic viscosity for gas turbine liquid fuels. It is common practice to arrange fuel dynamic viscosity as an input value to a performance computer program. Hence suitable values may be taken from Figures 7–22 and 7–23. If necessary the reader may “fit” a polynomial to the data presented in these charts to embed the properties within the computer program.

Liquid fuels are not “pumpable” with a kinematic viscosity of less than 1 cSt, and atomization will be unsatisfactory above 10 cSt. Figure 7–22 shows that kerosenes generally have a viscosity of less than 10 cSt even at –50°C and hence are always acceptable for atomization. However, diesels do exceed the limit and hence fuel heating is required if its temperature is allowed to fall below the threshold level.

None of the above effects are of importance for gas fuel as its viscosity is an order of magnitude lower; hence a database is not provided here.

CP of Fuels

For fully rigorous modeling fuel CP is required to calculate enthalpy input of the fuel prior to combustion. Approximate levels of CP for the primary gas turbine fuels are:

Kerosene to Diesel

The gas properties of the combustion products of kerosene and diesel, as well as their lower heating values are very similar. Hence there is negligible change in performance when burning either of these fuels.

Liquid Fuel to Natural Gas

The CP of the combustion products of the natural gas is around 2% higher than for liquid fuels. This has a noticeable impact upon performance increasing both the fuel energy flow required for a given combustor temperature rise, and the shaft power output. The net effect is an improvement in SFC. The exact change in performance depends upon the engine cycle but at constant SOT the changes when burning gas relative to a liquid fuel are in the range:

Oil Types

Oils used for gas turbine engines fall into two major categories: mineral oils and synthetic oils. Mineral oils result from the refining of crude oil. Usually anti-oxidant and anti-corrosive additives are employed for gas turbine use. Synthetic oils are ester based and their cost is approximately 10 times that of mineral oils.

Synthetic oils are used almost exclusively in gas generators due to their vastly higher auto-ignition temperature. Mineral oils are usually used for industrial power turbines where bearing chamber temperatures are sufficiently low. This is particularly true where journal as opposed to ball bearings are used, because the former requires approximately 20 times the oil flow to dissipate the heat resulting from significantly higher friction, hence the cost of using synthetic oil is prohibitive. Journal bearings are often required in large power turbines as they are capable of reacting far higher thrust loads.

Density

Formula F7.10 enables oil density for a typical synthetic oil with a density of 1100 kg/m³ at 15°C to be calculated as a function of oil temperature. The same formula may be factored to give the typical density of a mineral oil at 15°C, and then used for other oil temperatures. Typical values for oil density at 15°C are:

Viscosity, Kinematic and Dynamic

Kinematic and dynamic viscosity were defined earlier. Oil viscosity is important for calculating bearing and gearbox losses.

Figure 7–23 presents kinematic viscosity versus temperature for oils typically used in gas turbines. Light mineral oils are at the lower end of the band, synthetic oils in the middle and medium mineral oils at the top.

The formulae presented provide sufficient accuracy for all performance calculations.

Market Demand⁷

Contract awards for the last five years for fossil fuel-fired power plants (above 50 MW) averaged 63 GW per year, sales are forecasted to average about 67 GW per year over the period 1999–2004. From a regional viewpoint, the market in the Asia Pacific Basin is declining, while showing a moderate growth in Europe (starting from a low level) and strong growth in North America. This is reflected in the growth of the 60 Hz market. In the past, the proportion of power plants with gas turbines, that is, open cycle and closed cycle (combining gas and steam turbines), was around 54% of total fossil-fuel-fired power plants. For the next five years, this proportion is expected to rise to 63%. For open- and closed-cycle power plants, this means that sales are expected to grow from 34 GW per annum to 42 GW per annum. This growth is predominantly due to market developments in North America.

In comparison with coal-fired power plants, open- and closed-cycle power plants are characterized by lower investment costs. On the other hand, fuel costs (i.e., fuel price and plant efficiency) play a relatively more important role. Therefore, gas pricing is a very important issue, which ranges presently from about $2.0/GJ to $4.5/GJ, with the North American prices being at the lower end of the range. Thus, particularly in North America, open- and closed-cycle power plants can be particularly attractive investments.

Technology, Performance, Environment, and Design Trends

Over the last 50 years, the efficiencies of gas turbines went up from 33–38% (LHV) and the efficiencies of combined-cycle plants from 48–58%. At present, there are a rather small number of global manufacturers and suppliers offering GT products covering a wide range of power output. The products of the different suppliers can be grouped in clusters by comparable output. A rough classification is small (S), that is, up to 50 MW; medium (M), above 50 MW; and large (L), which is the class of the largest, high-performing heavy-duty gas turbines in the market with a power output above 200 MW. For all major manufacturers, the small and medium classes include proven existing GT models with lower efficiencies and operating temperatures but with robust and reliable performance; the large classes are confined to high-performance machines.

The latest gas turbine technology also reflects the expertise of aero-jet engine manufacturers. As a result, using pressure ratios of around 17, performance features have been achieved for heat rates in the range of 9400–9800 kJ/kWh and for efficiencies up to 38%. These performance improvements have been achieved through improved compressor design; higher turbine inlet temperature (approximately 1250°C ISO); advanced blading (design, materials, coating); optimal design of the turbine blade cooling system, with or without the need for external coolers; improved cooling of the combustion system sections (chamber, canes, transitions); and less pressure drop in the turbine exhaust.

Less than 25-ppm NO_x emissions can be expected from these machines. This compares with the maximum allowable levels of 60 ppm for GTs fired with natural gas as set out in the World Bank’s “Thermal Power-Guidelines for New Plants.” Because of better materials (alloys and coatings), with a higher resistance to high temperature oxidation and corrosion, combined with better cooling techniques, gas turbines are expected to reach a unit capacity of around 250–350 MW and GT efficiencies of 40% over the next years. In conjunction with optimized main steam parameters and improved plant component efficiencies, this in turn will increase the CC plant capacity into the 375–500 MW range, with CC efficiencies around 60%. (See also Figure 7–24.)

From a plant aspect, plant construction time has been shortened by the concept of predesigning, introduction of modules, and advanced project management and logistics. Inter alia, this was effected by simplification of the electrical and I&C systems as well as the foundation and overall civil works construction requirements. In addition, plant operation has been simplified, for example, by allowing full 100% steam bypass operation in the event of steam turbine malfunctions.

Cost Trends

Typical combined-cycle blocks are roughly in the range of 50–500 MW and costs range from $500–750/kW.

Fuels Rich in Hydrogen

Hydrogen (H₂) characteristics differ substantially from those of natural gas. H₂ exhibits low ignition energy, low molecular weight and density, and high flame temperature and propagation speed. Since H₂ has the ability to diffuse through and embrittle materials, the gas turbine itself and ancillary systems must allow for this fact.

For H₂-rich fuels, the combustors are particularly susceptible to premature ignition. Their design must avoid flashback by accelerating the fuel flow through burner passages while avoiding nozzle erosion caused by H₂’s high-speed flame. The diffusion combustion must use substantial amounts of steam cooling diluents for NO_x reduction because of the high flame temperature. Therefore, the combustion products have a high moisture content and a heightened heat transfer capability, which affect the life of hot-gas-path components. So the use of gaseous fuel with high H₂ content (above 50% volume) is currently limited to firing temperature typically found in older E-class gas turbine technology.

However, several combustion testing facilities are working on studying the behavior of full-scale combustor operation for F-class technology at firing temperatures and at actual operating pressures and flows. In several IGCC plants, fuel with more moderate H₂ content (between 20 and 30%) is combusted in diffusion mode at F-class firing temperature. Development of dry low NO_x combustors operating with higher than 10% H₂ content (the present maximum allowable concentration) will be a major breakthrough.

From both the end users’ and the EPC contractors’ perspective, the use of H₂-rich fuels presents many challenges, such as providing a different startup fuel capability, ensuring adequate purging systems to cope with much larger fuel handling systems volume (low BTU value), and dealing with stringent safety issues for leakage detection, explosion-proofing, ventilation, and so on. Special consideration should be given to materials of construction for pipes and valves as well as welded connections and sealing materials.

One EPC contractor’s (Bechtel) experience with medium H₂ content fuel is at Tampa Electric Company’s Polk Power Station Unit 1, an IGCC plant with a syngas fuel content of approximately 37% H₂ by volume at a lower heating value (LHV) of 200 BTU/ft³ (20% of the natural gas LHV).

Liquefied Natural Gas

To supply natural gas from the extraction fields to end users over large distances or oceans, the most common method is to liquefy the natural gas by refrigeration at –160°C (256°F) and then transport it in spherical tankers. At the receiving terminal, the liquefied natural gas (LNG) is unloaded and stored in insulated tanks. From the storage tank, the LNG is pumped at the desired pressure and vaporized for use in gas turbines. The combustion of vaporized LNG in a gas turbine is not much different than that of regular pipeline-quality natural gas in gaseous form. The use of LNG can be noted in many Asia-Pacific countries, where local natural gas resources are scarce.

It is expensive and therefore is used in countries such as Japan, which also has higher emissions standards than those in the rest of SE Asia.

Based on at least nine major projects, one major EPC contractor’s experience shows that integration of the LNG unloading, storage, and regasification facilities with the associated combined-cycle power plant yields substantial advantages to the owner. Sizable cost savings can be achieved by developing common auxiliary systems for cooling water and fire protection, by optimizing the site layout, and by integrating the unloading pier with power plant water intake and discharge structures.

A special feature of integration of the two facilities (LNG terminal and power plant) is the recovery of LNG cold energy from regasifying the LNG. In the past, the regasification was done by heating with seawater using open rack vaporizers without the heat. For the integrated system, one patented process uses a glycol-water mixture as the working fluid, which is cooled to 1°C in a closed-loop LNG vaporizer heat exchanger. The cold glycol water is then sent to the gas turbine to cool the gas turbine inlet air to 7°C. A fin-tube-type air-cooling coil is located downstream of the air filters. The return glycol water is pumped back to the LNG terminal, where it passes through a seawater heat exchanger to add supplemental heat as needed. Following the return of the warm glycol-water working fluid, the regasification process starts over again. Figure 7–26 illustrates the benefits of inlet cooling.

Liquefied Petroleum Gas

Liquefied petroleum gas (LPG) typically consists of butane and/or propane. LPG is a natural derivative of both natural gas and crude oil. Since LPG is composed mainly of heavier hydrocarbons, the gas turbine combustion system operates quite differently than for LNG. As the price for butane and propane varies seasonally, the receiving terminal should be designed to receive LPG in any combination of the two hydrocarbons, from 100% propane to 100% butane.

LPG is not widely used as a fuel for gas turbines because of its limited availability and the complexity of the fuel handling equipment. In terms of phase change properties at atmospheric pressure, propane is liquid at –42°C (44°F) and butane at 0°C (32°F). For use in gas turbines, the LPG is supplied at 90 bar (1300 psi); at this pressure, it can easily change phase between liquid and gas. One result of LPG’s volatility is that gas turbines can operate in only two modes, either “on” at full load or “off” at full speed, no load. With LPG supplied as 100% liquid at a lower pressure and with careful temperature control, gas turbine operation across the entire range of loads is possible. The limited market for LPG has kept manufacturers from developing dry low NO_x combustion systems for this fuel. The use of conventional diffusion burners necessitates the use of diluents to meet emissions limits, which affect the equipment life and increase the cost of the water treatment plant.

LPG vaporization can be also used to cool the air inlet to the gas turbine in the same way as LNG cold utilization. The economics of the inlet chilling systems, however, are quite different due to the diverse cryogenic properties of pentane and butane.

Today dimethyl-ether (DME) manufacture is commercialized. DME, which is interchangeable with LPG, can be manufactured from multifeedstocks such as coal, natural gas, and biomass in the same way as methanol. Table 7–3 lists properties of DME in comparison with butane and propane.

This multipurpose fuel is very attractive for developing countries (China, India) in permanent need of portable (bottled) or easily transportable fuels. OEMs (GE in particular) are looking at commercial offerings of DME-fired E- or F-class heavy-duty turbines. In Japan, the DME forum (JDF), formed in 2000, with 37 Japanese companies and two universities as members, works to ensure large-scale DME technology development. With low emissions and good thermal efficiency, DME could play an important role in specific power generation markets.

Fire detection and protection systems must be designed to allow for the fact that LPG and DME are heavier than air. Further, the fuel treatment system requires (1) fuel preheating to ensure that the fuel temperature is maintained above its dew point at all times and conditions and (2) proper handling and storage of the fuel.

Syngas and Other Low-Btu Gases

Gasification of coal and other low-grade fuels to produce syngas enables more efficient use of these fuels through IGCC technology. Over 40 projects worldwide (existing or in development) use heavy oil, petroleum coke, coal, biomass, and waste materials as fuel. The capital cost continues to drop as a result of technological advances and incorporation of lessons learned from existing facilities. Under the generic name of low-BTU gases, these fuels can vary significantly from one application to another and are highly dependent on the type of feedstock and gasification process. Table 7–4 summarizes some of the low-BTU gases based on hydrogen content and the range of LHV.

Conventional gas turbine combustion systems and associated fuel pretreatment apparatus cannot be used unmodified for these fuels. Since each fuel has different properties, analytical (mainly computational fluid dynamics; CFD) and experimental work is conducted before field implementation. Low-BTU gases on the combustion system necessitate full characterization of the combustor performance in test facilities. However, there is no substitute for actual field operational experience with each type of fuel. As a result, extended startup time is required to fine-tune the gas turbine fuel supply system. Table 7–4 illustrates the wide variety of low-BTU gases used in gas turbines.

The Wobbe Index (WI), a measurement of volumetric energy (calculated using the LHV and specific gravity of the fuel), indicates a combustor’s capability to burn different types of fuels using a unique fuel supply system. Usually, the maximum allowable WI variability is around ±5% of its value. Beyond this limit, multiple gases can be combusted in the same gas turbine only if supplied by independent fuel trains with different control valves, manifolds, and fuel nozzles, which add complexity and cost. Table 7–5 depicts the composition of low-BTU gases in comparison with natural gas. The LHV and WI are calculated as a percent of their absolute values for natural gas.

As shown in the Table 7–5, the LHV of syngas is lower than that of natural gas, requiring a considerable fuel mass flow increase to match the necessary heat input. When fully fired on typical syngas compositions, the gas turbine produces up to 25% more power than a conventional machine operating on natural gas, due to this increased mass flow.

In addition to the complexity of the combustion process related to the fuel composition, a major concern for gas turbines firing syngas is the significant amount of contaminants usually found in this type of fuel.

Deposits and Degradation

Certain deposits, especially from trace metals, affect gas turbine integrity and performance. In one particular application, Ni deposits concentrated on the tip of the burner forced a redesign of the burner. In many cases, the deposits plug the blades’ cooling holes and cause overheating as well as cracks in the material. Special coatings on the blades often are used to minimize the corrosion.

Performance Degradation

Another negative effect of deposits is the accelerated thermal performance degradation. Limits on the allowable degradation, recoverable and nonrecoverable from base load conditions can be provided by the OEMs. Operating below these limits will increase the risk of compressor surge. In the development phase of a project, the impact of the recoverable and nonrecoverable degradation, scheduled outages, and so on needs to be analyzed.

Light Distillates in Heavy-Duty Gas Turbines

Heavy-duty gas turbines are capable of burning a variety of liquid fuels ranging from light petroleum distillates to heavy residuals. The combustion process of light distillates such as naphtha, kerosene, and No. 2 fuel oil is well proven through millions of hours of operation. The real challenges for gas turbine design and operation come from the use of heavier ash-forming fuels. The combustion system capable must efficiently burn the fuel while maintaining emission limits and hardware integrity.

Residual Fuel

There is a glut of cheap residual fuel on the market. Countries like China are keen to use it. Even after the cost of transportation via tankers from the Middle East is added, it is still an attractive option for fast developing countries. Power generation is the largest industrial sector in China, which will grow to be the largest power producer in the world. A look at China’s environmental standards reveals that the Chinese are tightening their limits, but they still do not address CO₂ greenhouse gas emissions, other than aiming for maximized fuel efficiency. For all other emissions, like NO_x, if OEMs keep to best available technology (BAT), the environmental regulation targets are easily met.

The OEMs' response to their newly industrialized country (NIC) clients is predictable. When presented with a technical challenge, residual fuel or otherwise, they find a way to work with it. So, overall, residual fuel, with its resultant CO₂ emissions that exceed those formed with natural gas combustion, will continue to grow in terms of usage. Not just gas- and oil-poor countries use it. One of Siemens’s most successful residual fuel plants is in Mexico. Some OEMs are doing interesting work on CO₂ mitigation generally by postcombustion absorption, reinjection into rock formations (which they expect will hold the gas for some period of time), and similar initiatives. To this point, there has been no real breakthrough on this front.

Residual or “Bunker” Fuel Combustion⁹

Traditionally, older E-class technology gas turbines have been used for combusting crude oils, residue oils, and other heavy oils, since their firing temperatures are lower. Because the fuel is cheap, high efficiency is not an issue. However, the combustion of these fuels can prove extremely detrimental to gas turbine life due to corrosion, erosion, and fouling of the blades. Hot corrosion due to trace elements such as sodium (Na), potassium (K), lead (Pb), and particularly vanadium (V) in the fuel has a major impact on gas turbine performance and service life of hot gas path components.

Small amounts of V, naturally found in liquid fuels, create low-temperature-melting (1250°F) vanadium pentoxide (V₂O₅) ash. At typical gas turbine operation temperatures, these V ash deposits are in molten form and accelerate oxidation of turbine blades and vanes. The presence of other alkali metal contaminants in the fuel, Na and K salts, leads to formation of sodium sulfates (Na₂SO₄) (the sulfidation attack) through reaction with sulfur (S) in the fuel, resulting in pitting of the hot path components.

Fuel Treatment Summary

Prior to combustion in the gas turbine, the fuel has to be pretreated to reduce the concentration of contaminants to meet the stringent requirements set in Table 7–7. The pretreatment process involves chemical reactions and mechanical cleaning systems.

Additives for Residual Fuel

Na and K salts are soluble in water and can be removed by on-site fuel washing. Other methods of fuel treatment include the use of high-temperature corrosion additives. Effective additives are tailored to a specific fuel composition. The most common additive is magnesium (Mg). Additives containing Mg are used primarily to control vanadium oxidation by increasing the ash melting point, allowing the ash to remain in a solid state on gas turbine blades and vanes. The melting temperature of the new ash compound, which contains Mg, V, and oxygen (O), is much higher (around 1238°C or 2260°F) and dictates the maximum allowable firing temperature of the gas turbine. Other additives, based on chromium (Cr) and silicon (Si), are used mainly to combat Na- and K-based contaminants.

Mg-based additives are generically of three types: water soluble, oil soluble, and oil dispersible. The water-soluble type is the cheapest; however, it requires extensive mechanical handling. Oil-soluble additives are the easiest to handle, since they are self-dispersing on line and off line in fuel storage tanks. However, they are five times more expensive than the water-based additive. Magnesium sulfate (MgSO₄) is the most common water-soluble product using a 1:1 ratio for Mg:V. An example of oil-soluble product is KI 150 from Petrolite using a 3:1 ratio for Mg:V.

Startup and Shutdown

To ensure that the entire gas turbine fuel system is free from waxy fuel buildup or plugging, startup and shutdown must be done using No. 2 fuel oil. Transfer to crude can occur after the unit has been synchronized. The process should be carried out slowly to ensure heat-up of the piping, since crude needs to be preheated to 275°F to ensure that no frozen blocks are formed. On shutdown, No. 2 fuel oil is required to purge the system from any residues of crude.

Normal Operation

During normal operation, it is necessary to remove the ash formation on first-stage nozzles. The rate of deposits depends on type of fuel and ambient temperature. Because the amount of air moving through the turbine is larger at low ambient temperatures, the allowable plugging should not exceed 5–7%. At higher ambient temperatures, such as in the summer, the allowable plugging can exceed 10–12%. To reduce the amount of off-line water wash, some on-line wash could be done using an on-line turbine abrasive system and injecting particles at the first-stage nozzle.

Off-Line Cleaning

This process takes 10–12 hours from shutdown to start and comprises the steps shown in Table 7–8. Table 7–9 describes a gas turbine typical cleaning cycle. The concentration of contaminants determines the actual cleaning schedule.

Inspection Intervals

Inspection intervals are specific to each installation, determined by the condition of the hot-gas-path parts. In addition to contaminant corrosion, fuel containing heavier hydrocarbons has a higher H₂ content, which releases a higher amount of thermal energy during combustion. Higher combustion temperature leads to poorer hardware condition and shorter intervals between inspections. The number of equivalent operating hours (EOH), on which maintenance intervals are based, increases more rapidly based on fuel contaminant concentration. This can mean a factor of 1.5 for each operating hour.

A typical expected maintenance schedule follows in Table 7–10.

According to GE, several older gas turbines (7EA, Frame 5, and Frame 6) have accumulated several millions of hours of operation using “light Arabian crude” with V concentration of 10 to 30 ppm. The fuel is supplied directly from the refinery and then washed on site. An Mg additive is added at a ratio of 3:1 Mg:V. The units start and stop on crude. GE reports that after 1 million hours of operation, no significant hot path damage was found. Note that the older machines operate at much lower firing temperature (949°C or 1740°F) than modern gas turbines with a firing temperature of 1343°C or 2450°F or higher.

Success with Nonconventional Fuels¹⁰

Siemens’ experience in this arena demonstrates additional success with nonconventional fuel. Table 7–11 provides a sample of naphtha- and heavy oil-fired power plants in operation and in the planning stage. As this figure shows, some plants (e.g., Kot Addu and Valladolid) accumulated 30–60,000 hrs of successful operation over five years or more.

As a result of the impacts of deregulation, environmental standards as well as usage as secondary fuels, up to 25% of the gas turbine output (in MWs) in the next 10 years could involve non-natural gas fuels. Naphtha is expected to be the dominant fuel in this category with a few thousand MW of new capacity being added every year.

Technology, Performance, Environment, and Demand Trends

Figure 7–28 compares the supply flows, emissions, and by-products of different 600 MW-class plants. With reference to environmental emissions (including GHG), it may be seen that IGCC plants compare favorably with pulverized-coal-fired steam power plants.

Depending on the degree of integration between the gas turbine and the air separation unit (ASU), either standard gas turbine/compressor configurations can be applied, or limited modifications are required, to compensate for the mismatch between turbine and compressor mass flows, which results from the application of gases with low heating values. Three principal options are available. The selection of the appropriate air and nitrogen integration concept depends on a number of factors to be considered on a case-by-case basis. A summary of the important criteria is provided in Figure 7–29.

Figure 7–30 sets out the principal criteria for selection of the different IGCC integration concepts. The denominated “fully integrated approach,” which was selected for the European coal-based demonstration plants, resulted in the highest efficiency potential, but it is to be considered as highly sophisticated from the operational point of view. Nevertheless, after some initial operational problems, the Buggenum IGCC facility has demonstrated that this type of plant can be operated successfully and with the required availability.

The nonintegrated concept with a completely independent ASU is expected to have advantages concerning simplification of plant operation and possibly in achievable availability. On the other hand, the loss in overall IGCC net plant efficiency compared with the fully integrated concept amounts to 1.5–2.5%. This concept therefore is primarily of interest for application in cases where efficiency is not the key factor (e.g., for the gasification of refinery residues).

The concept with partial air-side integration is an interesting compromise solution, with an only moderate loss in efficiency but improved plant flexibility compared with the fully integrated concept.

In summary, this OEM notes that

Fuel Properties that Affect Fuel Treatment

Liquid Fuel

Liquid petroleum fuels fall into two basic classifications: true distillate and ash-forming oils. True distillate oils are most frequently the lighter oils such as naphtha, kerosene, and Number 2 distillate oil. This class of fuels is clean burning and behaves in a manner similar to natural gas as far as the internal portion of a gas turbine is concerned.

Naphtha, NGL, kerosene, and Light Number 1 distillate oil have little or no lubricity. Special handling equipment will be required in order to prevent damage to the peripheral equipment such as the main fuel pump, fuel nozzle plunger, flow dividers, etc.

Ash-forming oils are normally the heavier oils, such as blends, crude, and residual oils. Since they usually contain trace metal contaminants, fuel treatment is required to remove or modify the effects of harmful constituents prior to combustion in the power plant.

The degree of equipment sophistication, plant equipment investment and operating cost required for this treatment are tied directly to key physical and chemical properties of the fuel. Physically, the parameters of interest include specific gravity, viscosity, flash point, pour point, and wax content. Chemically, the prime interests are the amounts of corrosive trace metals present, the overall fuel washability, and compatibility of one fuel with another. The latter is particularly significant for users who obtain fuel from multiple sources (Figure 7–31).

Stability of Fuels

Future fuels will be of considerably poorer quality in the years to come. The major factor of such oils may be their incompatibility when oils from different refineries are used.

Viscosity Breaking (Visbreaking—Thermal Cracking)

Visbreaking is a simple process in which the very viscous residues from the vacuum distillation or even the atmospheric residues are subjected to thermal treatment at 450–490°C and 17 bars pressure. This results in the formation of low molecular weight, low-viscosity hydrocarbons that considerably reduce the residual viscosity.

Because the obtained residual is occasionally too viscous, it is diluted with cracked gas oil. Cracked gas oil is used as a diluent in order to reduce the tendency to sediment formation (the separating out of asphaltenes). Because of the high asphaltene content, the question of compatibility with other heavy fuel oils is particularly important (formation of excessive sludge in the centrifuge).

From the 100% stock at 450–490°C and 17 bars that enters the visbreaker, 66–87% of it is a residual fuel with a viscosity of up to 1000 cSt at 50°C.

Note: Based on the new distillation processes where sludge and catalytic fines are produced, no electrostatic desalter is recommended as a fuel treatment system for gas turbine fuels.

Liquid Fuel Treatment

In the case of a true light distillate oil, treatment is required primarily to capture dirt and shipment contaminants that have not been removed by settling in the fuel storage tanks. Undetected and untreated, a single shipment of contaminated fuel can cause substantial damage to a power plant. Filters are the most commonly used device although centrifuges and sometimes other desalting methods may also be required to remove sodium if present as a result of seawater contamination.

Treatment and conditioning systems for ash-forming fuels are more complex because of the processing steps required to remove or control their characteristic trace metal contaminants. Particular attention must be paid to sodium, potassium, lead, and vanadium. Complete fuel treatment includes:

Lead is also corrosive to power plant components (turbines) and should be avoided. Since there is no economical technique for removing or providing an inhibitor, it is necessary to only purchase fuel with extremely low amounts of lead (less than 1 ppm).

Fuel Washing

Since heavy fuel oils are almost universally contaminated with salts and salt water, they normally require water washing to remove the water-soluble salts of sodium and potassium. Fuel washing involves mixing heated fuel with 5–10% of potable water along with a small amount (0.02%) of an emulsion-breaking fluid to aid in separation. The washing process is a two-stage extraction procedure with the extraction water flowing counter-current to the fuel. Each stage has a mixer to initiate contact of the washwater with the fuel, followed by a mechanism to separate the salt-laden water from the extracted fuel.

The number of extraction stages required depends on the salt level in the fuel to be processed and on certain properties of the fuel such as specific gravity and viscosity. Since the successful operation of the fuel-washing system depends on a working difference between the specific gravities of fuel and that of water, fuels with specific gravities above 0.991 may not be washable unless they are first blended with a compatible liquid that has a lower specific gravity such as distillate fuel.

Chemicals referred to as demulsifiers are added to the fuel to assist in breaking up difficult oil/water emulsions.

Water is added to the fuel oil and is subjected to controlled mixing in order to distribute the sodium-seeking water throughout the raw fuel. The mixture of water and oil flows to a separator that is a mechanical centrifuge. In the centrifuge, the oil and water are separated under centrifugal force, the cleaned oil flows into one circuit and the latent wash water with a high concentration of salts flows out from the centrifuge through another circuit.

The effluent water may require a treatment system to remove the entrained oil to comply with local codes.

For example, a customer specification of “no visible sheen” which is approximately 15 ppm of free oil in the water. Reduction of oil content below these values may require a complicated system.