Proceedings of the ASME TURBO EXPO 2003

Power for Land, Sea and Air

June 16-19, 2003 ,Atlanta, Georgia, USA

GT2003-38027
SOME EFFECTS OF SIZE ON THE PERFORMANCES OF SMALL GAS TURBINES.

C Rodgers
3010 N Arroyo Dr.
San Diego. Ca 92103. USA
crodgers @4dcomm.com

ABSTRACT A later offshoot was the military incentive to provide aircraft

auxiliary power units (APU’s) compatible with the prime
By the new millennia gas turbine technology standards the
propulsion turbojet logistics.
size of the first gas turbines of Von Ohain and Whittle would
Very small gas turbines for a variety of applications followed,
be considered small. Since those first pioneer achievements
but were never manufactured in the same quantities as prime
the sizes of gas turbines have diverged to unbelievable
propulsion engines. The smallest and perhaps the most
extremes. Large aircraft turbofans delivering the equivalent of
fascinating are turbojet engines for model aircraft, Gerendas,
150 megawatts, and research micro engines designed for 20
Pfister, 2000, with a size similar to a coca cola can. In the
watts. Microturbine generator sets rated from 2 to 200kW are
propulsion field, engines include jet fuel starters and
penetrating the market to satisfy a rapid expansion use of
expendable turbojets Rodgers, 1986, also a variety of small
electronic equipment. Tiny turbojets the size of a coca cola
generator sets have been deployed for defense applications
can are being flown in model aircraft applications.
over the years, Rodgers,1993.
Over the last five decades small aircraft Auxiliary Power
Shirt button sized gas turbines are now being researched
Units (APU's) have been produced in large quantities. Many
intended to develop output powers below 0.5kW at rotational
of these machines are sophisticated, having multiple spools
speeds in excess of 200 Krpm, where it is discussed that
and power plus bleed air capability. A major design
parasitic frictional drag and component heat transfer effects
requirement for these machines is light weight, and a compact
can significantly impact cycle performance.
package, hence the reliance on the non- recuperative approach.
These machines have high reliability, and utilize sophisticated
The demarcation zone between small and large gas turbines
controls to facilitate integration with the main propulsion
arbitrarily chosen in this treatise is rotational speeds of the
engines and the aircraft environmental system, with the result
order 100 Krpm, and above.
that by commercial standards their cost is relatively high.
Following deregulation of the Utility Industry, a need became
This resurgence of impetus in the small gas turbine, beyond
apparent for modern, and much smaller and more efficient
that witnessed some forty years ago for potential automobile
turbogenerators for the Distributed Generation (DG) market.
applications, fostered this timely review of the small gas
About a dozen companies in the USA, Europe and Japan have
turbine, and a re-address of the question, what are the effects
small gas turbine turbogenerator (now phrased microturbine)
of size and clearances gaps on the performances of small gas
programs in progress. Units in the 30-100 kW power range are
turbines?. The possible resolution of this question lies in
expected to be produced in significant quantities within the
autopsy of the many small gas turbine component design
next several years or so. The first generation of recuperated
constraints, aided by lessons learned in small engine
microturbines are based on the use of existing materials and
performance development, which are the major topics of this
proven technology, and typically have an efficiency close to
paper.
30 percent. An excellent example of an operating microturbine
embodying state-of-the-art technology is the Capstone 30 kW
1. INTRODUCTION.
compact turbogenerator, Graig, 1997. Over a thousand of
The successful development of the aircraft gas turbine turbojet these units have entered service for a variety of power
in World War II subsequently triggered a myriad of gas generation applications, and it has been demonstrated for
turbine engine configurations covering a host of applications vehicular use.
and power range. Predominant amongst these was the efforts
to develop a competitive small automobile gas turbine
competitive to the compression ignition reciprocating engine.

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 A very small turbogenerator that was introduced for NOMENCLATURE.
commercial use was the Nissan Micro Gas Turbine rated at 2.6 A Area
kW, Nakajima, et.al., 1995. The package envelope for this APU Auxiliary Power Unit
microturbine was similar in size to a picnic ice chest. C Compressor
Large gas turbines have the advantage of both economy of CFS Volume Flow
performance and scale. The reverse is true for very small gas D Diameter
turbines. Small engines have significantly lower aerodynamic EGT Turbine Exhaust Gas Temperature
efficiencies. This is a result of smaller blade heights, Reynolds g Gravity
number effects, tip clearance effects, manufacturing H Head
tolerances, surface finish and engine-to engine variation all of ICR Intercooled recuperative
which adversely affect efficiency. Also their geometries make k Specific heat ratio
blade cooling very difficult, and thus advances in turbine inlet kW Kilowatt
temperature are essentially dependent on materials technology. Krpm Thousands of rpm
This poses a second question, is there a miniaturization limit Mu DeLaval number = U2/ (g kRT) 0.5
to small gas turbines?. An attempt to resolve of this question is N Rotational Speed
addressed in this paper, yet unfortunately the term small gas Ns Specific Speed =ω CFS0.5/(gHad)0.75
turbine has a diverse connotation, liberally applied, Leyes, and P Pressure
Fleming, 1999, to gas turbines ranging in equivalent power
q Work factor = ∆H/ U 2
outputs from 30-3000 hp.
R Radial, or Pressure Ratio
A proposed yardstick, Penny, 1963, delineating the large gas
RWC Compressor Normalized Inlet Flow Parameter
turbine from the small, was classification by the first stage
turbine nozzle area, somewhat equivalent to piston engine =(Wc √T1 /A 1 P1 )/ (Wc √T1 /A 1 P1 )crit
displacement, as the three basic parameters effecting output RWT Turbine Normalized Inlet Flow Parameter
power flow, pressure, and turbine inlet temperature, all in = (Wt √T1 /A n P1 )/ (Wt √T1 /A n P1 )crit
combination dictate the nozzle area. Miniaturization imposes T Temperature
constraints other than thermodynamic however, amongst TIT Turbine Inlet Temperature
which are: U Tip Speed
• Component fabrication and machining limitations V0 Spouting Velocity = √2g Had
• Ultra high rotational speed bearing design W Airflow
∆ Difference
• Non -scalar elements such as controls and protection
instrumentation η Efficiency
Lower Reynolds numbers do not inherently present a physical ω Angular Velocity
barrie r providing that the design major cycle performance Subscripts
parameters still provide net output power (or at least self- ad Adiabatic
sustaining conditions. Partial corroboration of this is the c Compressor
commercial manufacture of light vehicle turbochargers with d Diffuser
rotational speeds of 220krpm, Hayashi et,al., 1983. crit Sonic conditions
Harnessing silicon micro-fabrication techniques, Ashley, 1 Inlet
1998, is being used to develop tiny so called Mesoscopic gas t Turbine, or total
turbines with power ratings of less than 1 kW, which further
confuses the classification between small and big. The 2. SMALL GAS TURBINE DESIGN CONSIDERATIONS.
demarcation zone between small and big gas turbines The performance characteristics of small gas turbines depends
arbitrarily chosen in this treatise is rotational speeds skirting upon the choice of cycle configuration, which in turn is largely
100 Krpm, and above. Some justification for this choice arises dependent upon the application.
from the recommended antifriction bearing design life “DN” The various cycle schematic options shown on Fig 1 by no
criteria of approximately means cover the whole gambit, and are presented as
1.2 E 06 (mm, rpm), where speeds above 100krpm would limit representative of what is either in production, in development,
bearing bore and shaft diameters to no higher than 12mm. or has been researched in the past.
Less robust smaller antifriction bearings are of course The thermodynamic performances at design point conditions
available, so too is the option to use air bearings, as will be are basically a function of three parameters;
dis cussed later. • Peak cycle temperature, or turbine inlet temperature,
TIT
• Pressure ratio.
• Component efficiencies.

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 The TIT is essentially determined by the turbine rotor alloy Representative thermodynamic performances various gas
stress rupture and low cycle fatigue strengths, duty cycle, and turbine cycles including the turbo-compounded diesel are
rotor cooling options. The pressure ratio is dictated by the depicted on Fig 3,where thermal efficiency is shown versus
compressor type and material. specific power with TIT, and pressure ratio, as parameters. It
The component efficiencies are related to the rotor sizes, is apparent that more complex cycles such as the intercooled
aerodynamic excellence, and clearance gaps. recuperative cycle may provide higher thermal efficiencies,
In contrast to larger industrial gas turbines the thermodynamic but incur increased cost and durability issues.
performances of small gas turbines are therefore size (output The small gas turbine designer is then shackled with the
power) dependent, and scalar techniques so widely employed dichotomy of maximizing performance and durability all
in the design of larger engines can no longer be strictly confined within a compact envelope at a competitive
adhered to. An example of scalar relationships for larger manufacturing cost. The resolution of this dichotomy centers
engines with single stage centrifugal compressors is shown on upon component number reduction and engine flowpath
Fig 2, showing relative changes in weight, speed, flow, torque, simplification.
inertia and acceleration time versus power ratio. Note that The influence of size on the design of small gas turbine
even on Fig2 some departure from true scale is portrayed, as engines and components encompasses a variety of
an examination of actual engine weights in a 300-50kW class considerations and compromises, which may become more
critical in subsequent manufacturing, assembly, and test
gas turbine revealed that the weight was departing from the
qualification phases than possibly larger gas turbines. Some of
cube/square 1.5 power law, to a 1.2 power, presumably due to
these considerations are discussed as follows.
fabrication techniques and cost constraints.

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 2.1 Engine Flowpath. 2.2 Compressor.
A collage of small gas turbine flowpaths is shown on Fig 4, Choice of compressor type and pressure ratio is a major design
representative of existing engines but not inclusive of every consideration. Small gas turbines have been designed with
single flowpath that has been developed and tested. pressure ratios ranging from 2.5 to 8.0 with both single stage
Examination of these flowpaths and component arrangements centrifugal, multistage axial compressors, and combinations
tend to substantiate the emphasis on reduced component thereof. The centrifugal is the least sensitive to clearance
number. losses, and is furthermore capable of exhibiting wide surge
margins with high inlet flow distortion tolerance.
The simplicity and reduced cost features of the single stage
centrifugal compressor are ideal assets for low cost gas
turbines with output powers from 10 to 300 kW.

The attainable efficiencies of single stage centrifugal

compressors, with ambient air suction, and are largely
dependent upon four parameters.

• Inlet Specific Speed Nsc

• Impeller tip diameter (size)
• Blade backsweep
• Inducer tip Mach number.

Although Nsc serves as an initial design guideline, choice of

Nsc “a-priori” is no unconditional guarantee of peak attainable
performance, thus good design, manufacturing, and test
development procedures are still essential. Furthermore the
compressor is only one engine component and the turbine
Fig2. Engine Power Scalars design is equally important, particularly for a shaft power
engine type where the turbine may provide up to 30 % more
Although simplifying design through component reduction can power than the compressor. In small turbojet applications
be construed as an overstatement of probable reliability, the requiring maximum thrust per frontal area both compressor
lower parts count of the single shaft unit is at least conducive and turbine efficiency may be compromised in favor of
towards lower cost and improved durability. The combination airflow swallowing capacity.
of a radial comp ressor and turbine mounted back-to-back, in
single shaft form, has found wide acceptance particularly in Specific speed too involves a choice of rotational speed, which
small microturbines, APU’s, and expendable turbojets. influences shaft and bearing dynamic design, a critical concern
with rotational speeds in excess of 100krpm. The influence of
The mechanical design approach of low pressure ratio, lower specific speed of both the compressor and turbine on engine
rotational speed, multiple stages for small gas turbines has design and engine miniaturization will be discussed later.
been previously studied, Wilson and Korakianitis, 1998, but
has found limited application, at least within the relevant The typical efficiency levels of small single stage centrifugal
speed range. Lower rotational speeds induce lower stresses compressors with abradable shrouds, and vaned diffusers,
and cyclic loads on the rotational parts, permitting simplified Rodgers,1991, are shown on Fig 5 as a function of specific
lower cost bearing and shaft dynamic designs. speed and flow, for pressure ratios of 3.5 to 4.0.
With exception of the automobile gas turbine, most heat Flows less than 0.1 pps are observed to reduce efficiency
exchanged small gas turbines have embodied stationary below 70%, as corroborated by a US Army funded study, and
recuperators of the wrap-around annular, or modular add-on compressor component testing of a 1.5/3 kW unit by
type. Airesearch. This compressor, Faehn, 1975, was tested at a 154
krpm, delivering a stagnation pressure ratio of 4.0, flow of
0.16 pps, and adiabatic efficiency of 72% t-t.

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 A test data point performance of this compressor with a particularly prone to compressor heating. Ten (10) deg F of
clearance of .005 inch is superimposed on Fig 6, for reference inlet heating was predicted for a 200KW monorotor engine
purposes. with a compressor Nsc of 0.8, and that for the same cycle
Transonic inducer conditions are approached as pressure ratios conditions the inlet heating would increase inversely
rise above 4.0, provoking shock losses , which can be defrayed proportional to Nsc2 . Halving the engine design point speed
by adopting razor thin blade leading edges, combined with would therefore increase the inlet heating to 40F, and
reduced suction surface turning. Such small blades may potentially reduce power by 20%.
however become impractical to manufacture, and accordingly it
is prudent to select lower design pressure ratios as size is
diminished.
The effect of axial shroud clearance on compressor efficiency
can be correlated with the clearance gap/tip blade height,
indicating that a gap ratio of 10% can decrease efficiency 4%
points. Clearance losses can be mitigated utilizing abradable
shrouds, permitting the impeller blades to machine their own
minimum operating clearance, (providing the blades are robust
enough). The use of shrouded impellers does not entirely
mitigate clearance effects, as a seal must be provided to throttle
flow recirculation back to the entry, in addition to which
rotating shrouds complicate impeller manufacturing.
Fig 5. Compressor Efficiency vis Nsc
Heat transfer from the hot turbine to the “cold”compressor can
The US Army 1.5/3KW gas turbine centrifugal compressor
also be a problem, since gas turbine engine power lapse rate
map is shown in normalized format on Fig 6, where inlet flow
characteristic is such that the influence of 10 degree F inlet
parameter RWC is plotted versus pressure ratio and efficiency
temperature rise above ambient at the inducer inlet can reduce
with parameters of De Laval number Mu. The normalized map
engine power approximately 5%. Monorotor type compressor
is fairly universal for medium Nsc sweptback compressor
and turbine assemblies, Rodgers, 1978, are therefore
designs with vaned type diffusers, and may therefore be used as

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 guideline in preliminary design studies. State-of-the-art total-to-ambient efficiency levels (i.e., with
exhaust diffuser) for small radial and axial turbines are shown
The inlet flow parameter RWC is based upon the inducer inlet on Fig 7, as a function of specific speed (Nst) with airflow as a
unblocked annulus area, whereas actual choke flow is parameter.
controlled by either, the inducer, or the diffuser throat area. These levels are representative of metallic turbines with typical
Higher Nsc designs typical of small turbochargers exhibit blade finenesses, with minimum clearance gap/blade height
maximum values of RWC approaching 0.8. ratios, of the order 5%, to prevent shroud contact under
transient operation. The effect of axial shroud clearance on
radial turbine efficiency, Rodgers,1997, can be correlated with
the clearance gap/tip blade height, indicating that a gap ratio of
10% can decrease efficiency 3% points. Axial turbines are
approximately three times more sensitive to clearance gap
blade height ratio.

For common compressor and turbine speeds, Nst is related to

Nsc by the relative compressor and turbine densities and
adiabatic heads. Typical specific speed ratios Nst/Nsc for
single stage centrifugal compressors are approximately 0.85 for
shaft power engines, and 1.05 for turbojets.
In comparison the design of microturbines with single stage
radial compressors and turbines is focused towards
optimization of thermal efficiency, thus rotational speed
2. 3 Turbine Types. selection becomes a compromise between both Nst and Nsc. In
Inward flow radial and mixed flow turbines have established a this respect it should be noted that the sensitivity of compressor
prominence in small turbomachines because of their simplicity, and turbine efficiencies on engine performance is not only on
low cost, relatively high performance, and ease of installation. the singular effect of component efficiency changes
Two predominant applications of these turbines are in small themselves, but also the result the efficiency changes may have
gas turbines and turbochargers, with airflows in the 5.0 to 0.01 upon other cycle variables. Two examples are, 1) increasing
pps range. compressor efficiency at a given speed will normally also result
in and increase in pressure ratio, and more important a
congruent increase in airflow for a given turbine nozzle area.
Alternatively, 2) increasing turbine efficiency can decrease
turbine exhaust temperature, which if utilized as the basis for
engine rating (e.g with a recuperator) can permit higher power
output than reflected by the efficiency increase alone.

Fig 7. Turbine Efficiency vis Nst.

Small gas turbines with turbine expansion ratios above 3.0

result in overloading the single stage axial turbine; higher
efficiencies can be obtained with two axial stages, or single
stage higher tip speed, radial or mixed flow turbines. Design Fig 8. Turbine Efficiency vis U/V0
studies have indicated that for the same overall turbine
efficiency the single stage radial can offer the lowest cost, but The flow controlling elements of the turbine stage are the nozzle
requires a larger diameter with more containment weight. throat, and rotor throat areas. Typical pressure/flow and speed
characteristics of radial and axial turbines are shown on Fig 8 in

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 terms of the relative flow parameter RWT and velocity ratio The above arguments may become flawed in actual engine
Ut/V0. The rotor exit throat area is determined by desire for both design, as in contrast high Nsc designs with higher rotational
low exit swirl and low exhaust energy, with proviso that the speeds, influence both shaft and bearing journal stiffness. At
blade root centrifugal stress, governed by AN 2 does not exceed speeds of the order 300 krpm up to 25% of engine power output
approximately 8.0 × 10 10 (in 2 , rpm) may be consumed in hydrodynamic air bearing drag, dependent
upon tip speed, axial thrust.
2.4 Bearings.
2.5 Structural and Material Constraints.
Small gas turbines currently operate with rotational speeds
from 60 to 150 krpm with both conventional antifriction and air The three major design parameters that link engine
bearings. Bearing “DN” values approach 1.5*106 . Such high performance to engine life and durability, are
rotational speeds imply small bearings, which are sensitive to (1) Turbine inlet temperature,
internal clearances and thus temperature environment, (2) Compressor pressure ratio
especially hot-end, rear-bearing locations. Adequate cooling (3) Rotational speed.
either from the compressor discharge or a lubrication supply is These three are also coupled to turbine efficiency via the
paramount to prevent a runaway thermal condition. Alternate turbine velocity ratio Ut/V0, and specific speed Nst.
design options to ameliorate this hot end bearing environment During preliminary engine design and performance
are: optimization phases, it then becomes mandatory to conduct a
trade-off between turbine rotor efficiency and life, and thus the
• Front end air inlet housing bearing capsule location relationship between rotor stresses, operating temperatures, and
with overhung rotor assembly. required tip speeds.
• Use of ceramic bearings with lower radial loads and To assess stress rupture and low cycle fatigue life, most of the
higher temperature capability. following information is essential:
• Throw-away fuel lubricated bearings, often used on
• Rotor geometry
expendable or model turbojets.
• Self contained grease packed or wick fed systems. • Rotor thermal and stress model
• Definition of typical operating cycle
The simple Petroff equation relates viscous fluid • Material Stress/Rupture, Low Cycle Fatigue
film bearing losses as a function of: properties.
Loss ƒ D3 *N2 * viscosity.
An initial choice of turbine tip speeds and TIT’s for current
For fixed specific speed and cycle conditions, the
super-alloys of 2000 fps (610m/s) radial, 1600 fps (490m/s)
scalar relationships can be applied.
axial, and 1800F (982C) provides a dependable baseline with
Output power α W, thus α D2 .
which to initiate these turbine life computations.
The ratio of Loss / Power α D3 *N2 / D 2 .
Thus the ratio bearing losses to output power α N. 3. SIZE EFFECTS ON ENGINE PERFORMANCE.

This would translate to relatively lower mechanical efficiency Upon contemplation regarding gas turbine min iaturization
with increasing speed. Surprisingly high bearing losses of the limits, it becomes apparent that engine size reduction becomes
order 1.0kW were quoted by Nakajima,1995, for a 2.6 kW impractical when the engine component efficiencies and cycle
microturbine, the equivalent of 72% mechanical efficiency. conditions can no longer provide useful power output, i.e., no
Air bearings require no lubrication or associated lubrication longer self sustain. Examination of the Brayton cycle at self-
sustaining conditions shows that the self-sustaining maximum
cooling system, plus minimal parasitic drag during starting. Air
bearings however possess low thrust bearing load capacity, and cycle temperature ratio TIT/T1 is a function of ;
are sensitive to thermal gradients, and shock loading under high
impact "g" accelerations. TIT/T1 ƒ ( ηc ηt ηm) ( Rc/Rt ) (Wt/Wc)
Air bearings too incur power losses, especially the thrust
bearing, which may be as large in diameter as the compressor This self-sustaining parameter is shown plotted on Fig 9
showing that typical turbine rotor superalloy temperature limits
impeller. The impeller exhibits almost constant tip speed ( N α
1/D), for a given pressure ratio and geometry. With a fixed (TIT/T1 ≅ 4.0 ) would be likely exceeded when the value of this
diameter ratio between the compressor and thrust bearing the function falls below 0.35. The relationship was found to be
absolute bearing loss then becomes proportional to diameter, surprisingly independent of pressure ratio.
which further implies that for a given engine design low Nsc With flow continuity, a mechanical efficiency of 95%, and a
selection leads to higher thrust bearing losses. pressure losses amounting to Rc/Rt = 0.95 the corresponding
minimum product of ηc ηt for simple cycle self sustain would
be 0.39, necessitating both the compressor and turbine
efficiencies to be above 62%. Recuperated cycles with higher

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 pressure drops exacerbate this condition. Ceramic materials
might of course permit higher TIT’s (TIT/T1> 5.0) and thus
somewhat lower component efficiencies.
The sensitivity of self sustain conditions to component flow and
thus, size can be assessed by superimposing the efficiency
versus flow trends of Figs 5 and 7 on Fig 9, thereby illustrating
that with current state-of-art small turbomachinery technology,
microturbines with design point flows approaching 0.001 pps
may challenge the fringes of self-sustaining operation. Note that
the MIT Micro Gas Turbine Project, Epstein 2000, was designed
with a flow of .0005 pps and TIT/T1= 5.5. With compressor Nsc
of 0.3 estimated compressor inlet heating, Rodgers, 1978, would
have been 300F!.
This projection was substantiated by additional cycle
Fig10. OptimumThermal Efficiency Regimes
performance analyses using a computational procedure
Rodgers,1997, addressing the effect of engine power output on
Note that as the engine size diminishes the speed islands also
performance, physical size, weight and cost. Commencing with
diminish which is a consequence of specific speed efficiency
an initial selection of power output, pressure ratio, TIT, and trends converging towards the self-sustaining domain and into
component type, with rotational speed as input parameter, the
an eventual black hole. Component efficiencies skirting the self-
procedure iterates upon component efficiencies via Reynolds
sustaining level may also be deleterious to engine acceleration
number and component specific speed relationships, to converge characteristics, necessitating over temperature past the
upon preliminary engine dimensions. An original motivation for
externally assisted torque cutoff to rated speed conditions. Over
deriving the computer procedure was to enable the project
temperature is the nemesis of small metallic turbine blade
engineer to conduct thermodynamic and economic trade studies heights and thicknesses. Blade thicknesses below 0.03 inch can
in the preliminary design phase allowing earlier identification of
readily vanish even at moderate TIT’s, if subject to high
a promising component candidate configurations.
combustor temperature pattern factors.
A huge domestic market potential is foreseeable for a 5kW
microturbine, but a 1.0kW unit would seem to have a limited
market perhaps other than special military applications and
logistic requirements.

Fig 9. Self Sustaining Cycle Conditions.

Fig11. Effect of Clearances on Output Power.
The procedure was applied to conduct preliminary design
analyses of a 5kW microturbine McDonald and Rodgers, 2001, In order to emphasize the importance of clearances on small gas
and subsequently exercised in this paper to examine the turbines the estimated effect of increasing the 5KW
anticipated performance of a 1.0 kW sized microturbine, with microturbine compressor and turbine clearances from .005 to
either metallic or ceramic hot end components. .015,and .015 to .025 inch, is shown on Fig 11 indicating a
The results of this exercise are shown on Fig 10, revealing that power loss of 12%, excluding the effects of throat area
optimum rotational speeds would increase to some 350 krpm variations. Thus tolerance and clearance control would
metallic, and 450 krpm ceramic, with decreasing thermal obviously be prime engine assembly stack-up and inspection
efficiency below 20%. criteria.

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 CONCLUSIONS. A 100 watt micro gas turbine, Isomura, et,al., 2002, is currently
being researched in Japan. Nevertheless it is anticipated that at
The resurgence of potential market applications of the small gas
power levels above 1.0 kW the small metallic gas turbine, and
turbine fostered a review of the effects of size on the
possibly eventually the small ceramic gas turbine could be
performances of small gas turbines, encompassing an
economically viable if mass produced with automobile
examination of small gas turbine engine design constraints and
manufacturing technology capabilities.
developmental lessons learned. Smallness was arbitrarily
This is as partially substantiated by the Nissan 2.6kW
defined as engines with rotational speeds exceeding 100 krpm
microturbine development program, and witnessed by small 10
which although methodically convenient, excludes practical
lbf, 160,000rpm turbojet kits, already available for a purchase
multistage engine flowpaths with equally lower power outputs
price of the order $1500.
but with lower rotational speeds, thus lower dynamic stresses,
and the feasibility of non-metallic materials.
A lessons learned approach addressing the many facets of
component size effects on small gas turbine performance and
A huge domestic market potential is foreseeable for a 5kW
viability was adopted. Lessons learned can impair vision, as
microturbine, but smaller units would seem to have a limited
history records of the 1940 National Academy Scientific panel
market perhaps other than special military applications and
that stated gas turbines would never fly, when indeed the
logistic requirements. Commercial market potential is uncertain
Hienkel 178 had already flown in 1939 powered by the Von
basically from both demand and profit margin considerations.
Ohain turbojet.
Analytical resolution of the question first focuses towards the
Shirt button sized gas turbines are being researched and
component efficiencies necessary to provide, at the very least,
optimistically will at least demonstrate self-sustaining
self-sustaining operation, difficult indeed for the early gas
operation, as to whether they can prove economically viable
turbine pioneers. Analytical studies herein suggest that units of
remains a trial of need versus cost.
1.0 kW could become feasible with the proviso that research
technology is spearheaded towards three critical elements:

(a) Stable consistent combustor operation on liquid fuel.

(b) Tolerable bearing viscous drag.
(c) Reducing turbine to compressor heat transfer.

These elements are conceivably more critical than the effects

of turbomachinery miniaturization on attainable compressor
and turbine efficiencies.

Current manufacturing technology tolerances for both metallic,

and ceramic castings, are projected to cause large performance
variations, beyond those already incurred in small
microturbines, pending newer technology fabrication
techniques and engine assembly procedures. Advanced
technology plasma layering and stereo lithography procedures
are in development but it may be premature to suggest these
techniques may significantly impact the tolerance variation
concerns. Layering techniques together with five-axis milling,
Childs and Noronha, 1997, are intimately coupled with the
trade offs between elapsed time and surface finish.

Alternative electrical energy sources, fuel cells, high energy

density batteries and thermo -electric generators, are being
researched which may prove both economically viable and
power density competitive with lower power sources in the 50
–500 watt bracket.

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 REFERENCES. Rodgers, C.,1991. “The Efficiencies of Single Stage
Centrifugal Compressors for Aircraft Applications”. ASME
Ashley. S., 1998 “Palm –Size Spy Planes”. ASME Mech Eng
91-GT-77.
Feb 1998.
Rodgers, C., 1993, “Small (10 - 200 kW) Turbogenerator
Craig, P., 1997, “The Capstone Turbogenerator as an
Design Considerations”, ASME IGTI- Vol.8, pp 535-542.
Alternative PowerSource”, SAE Paper 970292.
Rodgers, C.,1997. “Thermo - Economics of Small 50 kW
Childs, P.R.N., Noronha, M.B., 1997. “The Impact of
Turbogenerator”, ASME paper 97-GT-260.
Machining Techniques on Centrifugal compressor Impeller
Performance”. ASME 97-GT-456.
Wilson,G.W., Korakianitis,T.,1998. “The Design of High
Efficiency Turbomachinery and Gas Turbines”. Prentice
Epstein, A. K., et al, 2000. “Shirt -button-Sized Gas Turbines:
Hall.
The Engineering Challenges of Micro High Speed Rotating
Machinery”. ISROMAC-8 2000.

Faehn, D., 1975. US Army MERDC 1.5/3KW Gas Turbine

Generator Program. Ft Belvoir Va.

Gerendas, M, and Pfister, R., 2000, “Development of a Very

Small Aero Engine”, ASME Paper 2000-GT-0536.

Hayashi, M., Okazaki, Y., Tsujimura, H., 1983. “Development

of Mini-Turbochargers for Passenger Car”, Gas Turbine
Society of Japan. Tokyo 1983.

Isomura.K., et.al., 2002. “Development of Microturbocharger

and Microcombustor for a Three-dimensional Gas Turbine at
Microscale”, ASME GT-2002-30580.

Leyes, R.A.,Fleming, W.A., 1999. “The History of North

American Small Gas Turbine Aircraft Engines”, AIAA,
Smithsonian Inst.

McDonald, C.F., Rodgers,C.,2001. “The Ubiquitous Personal

Turbine (PT).. A Power Vision for the 21st Century”. ASME
2001- GT 100.

Nakajima, T., et.al., 1995, “The Development of the Micro Gas

Turbine Generator”, Paper Presented at Yokohama
International Gas Turbine Conference.

Nakazawa, N., et al.,1997. “Status of the Automotive Ceramic

Gas Turbine Development Program-Seven Years of Progress”,
ASME 97-GT-383.

Penny.N., 1963. “The Rover Case History of Small Gas

Turbines”. SAE Paper 63 – 634A.

Rodgers, C., 1978. “Design and Development of a Monorotor

Gas Turbine APU”. ASME 78-GT-2.

Rodgers, C., 1986, “A Jet Fuel Starter and Expendable

Turbojet”, ASME Paper 86-GT-1.

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