Micropower generation with microgasturbines: a ...

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Micropower generation with microgasturbines: a challenge J Peirs1∗ , T Waumans1 , P Vleugels1 , F Al-Bender1 , T Stevens1 , T Verstraete2 , S Stevens3 , R D’hulst3 , D Verstraete4 , P Fiorini5 , R Van den Braembussche2 , J Driesen3 , R Puers3 , P Hendrick6 , M Baelmans1 , and D Reynaerts1 1 Department of Mechanical Engineering, Katholieke Universiteit Leuven, Leuven, Belgium 2 von Karman Institute for Fluid Dynamics, Sint-Genesius-Rode, Belgium 3 Department of Electrical Engineering, Katholieke Universiteit Leuven, Leuven, Belgium 4 Royal Military Academy, Brussels, Belgium 5 IMEC, Leuven, Belgium 6 Université Libre de Bruxelles, Brussels, Belgium The manuscript was received on 11 August 2006 and was accepted after revision for publication on 12 December 2006. DOI: 10.1243/0954406JMES472

Abstract: This paper describes the development of a microgasturbine with a rotor diameter of 20 mm. The target electrical power output lies around 1 kW. The total system fits in a cylinder with a diameter of 95 mm and a length of 120 mm. The system contains the same components as a large gasturbine generator: compressor, recuperator, combustion chamber, turbine, and electrical generator. Major challenges are the high rotational speed (500 000 r/min), high turbine inlet temperature (1200 K), and the efficiency of the components. Because of the small dimensions, the flow through compressor and turbine is characterized by relatively low Reynolds numbers. The higher flow losses and inherently lower efficiency require a higher blade tip speed (524 m/s) than for large turbines (300–400 m/s). To minimize wear and frictional losses, the rotor is mounted on aerodynamic bearings. To withstand the high centrifugal stresses, a high-strength steel is used for compressor and shaft. The turbine is made of a Si3 N4 –TiN ceramic composite to withstand the combination of elevated stress and temperature. Keywords: powerMEMS, microturbine, microgenerator 1

INTRODUCTION

The demand for portable electrical power is steadily increasing. Despite the recent advances in battery technology, the gap between required and available power is still increasing. Fuels have much higher energy densities (J/kg or J/L) than batteries [1] and therefore can bridge the gap, but only on the condition that efficient miniaturized devices exist that convert the chemical energy of fuel to electrical energy. Such power converters can be based on a wide range of operating principles: fuel cells,

∗ Corresponding

author: Department of Mechanical Engineering,

Katholieke Universiteit Leuven, Celestijnenlaan 300B, Leuven B-3001, Belgium. email: [email protected]

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thermo-electric devices [2, 3], thermophotovoltaic devices [4], external combustion engines [5] such as Stirling engines [6, 7], internal combustion engines [8–10], wave disk engines [11], gasturbines [12–21], and steam turbines [22]. In analogy to large power systems, microfuel cells are expected to offer the highest efficiency, while microgasturbines are expected to offer the highest power density (power generated per unit volume). As the energy density of the total system – generator plus tank – has to be considered, both efficiency and power density of the energy conversion device are important to take advantage of the high energy density of the fuel. In the literature, two different turbine size ranges can be found: a microrange (MIT [12], Tohoku University [13], Onera [14], National University of Singapore [15]) and a mesoscopic range (Stanford Proc. IMechE Vol. 221 Part C: J. Mechanical Engineering Science

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[16], University of Tokyo [17], Tohoku University [18], ETH Zurich [19], Belgium [20]). The microrange uses silicon technology, enabling the production of very small components with high accuracy. The main drawbacks of this technology are the limited wafer thickness resulting in pancake rotor designs, short two-dimensional blades, and the limited choice of materials. The goal for the power output is in the range of 1–10 W of electrical power, with application in small portable electronic devices, wireless sensors, consumer electronics, etc. For mesoscopic devices there is more design freedom, as a wide range of production methods is available: micromilling [17, 18], micro-EDM [21], and micromoulding [16]. These production techniques allow the production of three-dimensional blade geometries in a wide range of materials (steel, nickel alloys, ceramics). The overall unit size envisaged is in the range of 1 dm3 with an electrical power output of 100 W–1 kW. These devices find application as portable power generator in civil and military applications, in unmanned aerial vehicles, as auxiliary power unit, to drive power tools, etc. The microgasturbine discussed in this paper fits into the upper power segment aiming at an electrical power output of about 1 kW. The way to develop such a microgasturbine is not simply to scale down a large gasturbine. All components have to be redesigned and often new concepts have to be used. Major problems are the high rotational speed (>500 000 r/min) and temperature (>1200 K), and the efficiency of the components. The current paper attempts to give an overview of the project as a whole, and therefore all major components and design issues are covered. The following two sections discuss the general layout and thermodynamic cycle. The next sections discuss the development of each component: combustor, bearings, rotor dynamics, compressor and turbine, recuperator, and generator. Finally, fabrication of the impellers is discussed.

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GENERAL LAYOUT

Figure 1 shows the general layout of the microturbine generator. The system basically consists of a compressor, recuperator, combustion chamber, turbine, and electrical generator. In total it has a diameter of around 95 mm and a length of 120 mm. The compressor and turbine impellers are 20 mm in diameter. The rotor takes only a small part of the total volume, in contrast to large gasturbines. This is typical for microturbines where the relative combustion chamber size has to be increased to get sufficient time for combustion. Similar effects can be noticed for the recuperator. In order to accomodate the relatively large volume of Proc. IMechE Vol. 221 Part C: J. Mechanical Engineering Science

Fig. 1

Microturbine layout

both the combustion chamber and the recuperator in a compact way, an annular design was chosen for both components. As a consequence of the adopted layout, the hottest part – the combustion chamber – is enclosed by the recuperator on the outside and by the exhaust diffuser on the inside. This allows for recycling of heat losses from the combustion chamber. The generator is located on the left side, far from the hot parts, as the generator magnets demagnetize at high temperatures. In addition, the inlet air is led through channels in the generator stator for additional cooling. An exhaust diffuser is added to create a subambient pressure at the turbine exit, such that more power can be extracted. Generator, compressor, and turbine are mounted on a single shaft. This choice creates several problems. 1. The generator has to operate at the same high speed as the turbine, subjecting the magnetic rotor material (magnets and/or iron) to high centrifugal loads and increasing the Eddy current losses and skin effects. 2. Heat flows through the shaft towards the generator where high temperatures could lead to demagnetization. 3. The length of the shaft results in low eigenfrequencies for bending (8–9 kHz) and torsion (∼17.5 kHz), compared to the rotational frequency (8.3 kHz). These problems could be solved with a two-shaft design: one high-speed shaft with compressor and high-pressure turbine and a second low-speed shaft with a low-pressure turbine and the generator (similar to the design of ETH Zurich [19]). The low-pressure turbine and generator operate at a lower temperature JMES472 © IMechE 2007

Micropower generation with microgasturbines

and speed than in the case of a single turbine solution. Both shafts are also shorter than a single shaft, resulting in considerably higher bending and torsion eigenfrequencies. Besides this the two-shaft design has two major shortcomings. 1. The generator cannot be used as a startup motor because it is not connected to the compressor and therefore cannot generate the pressure required to start the Brayton cycle. 2. A two-shaft turbine is difficult to control: both shafts have to be controlled in speed, one by the amount of fuel injected and the other by the generator load. Interaction between both shafts and the thermal time constants of the combustion chamber and recuperator complicate control. In contrast, a single-shaft turbine can be easily controlled by the generator with minimal time delay. Another possibility is to couple the two shafts with a flexible coupling. However, to have two dynamically independent shafts the coupling should be more than an order of magnitude more compliant than the aerodynamic bearings (which have already a stiffness below 0.1 N/μm). If not, both shafts will act as a single shaft with an increased flexibility in the middle, which is even worse than a single rigid shaft. On the other hand, such coupling has to be stiff and strong with respect to torsion, and be able to resist the centrifugal forces. No suitable design was found that meets these specifications. The conclusion is that the two-shaft design introduces more problems than it solves. Therefore, the singleshaft solution is preferred, and special attention will be paid to maximize the natural frequencies. 3 THERMODYNAMIC CYCLE AND EFFICIENCY The main difference between a small and large gasturbine is the amount of fluid submitted to an almost unchanged thermodynamic cycle. Hence the velocity

Fig. 2

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and pressure levels remain the same when scaling down a gasturbine; only the dimensions are smaller. The work exchange between compressor or turbine and fluid is proportional to the peripheral speed, such that the rotational speed should scale inversely proportional to the diameter. The performance of the turbo-machinery components remains unchanged as long as the following non-dimensional parameters are maintained. 1. Reynolds number, relative roughness of the surfaces, and rotational number (RoDh ): these have a direct impact on friction and hence on losses. 2. Mach number: characterizes compressibility effects. The Mach number is conserved for the same gas characteristics and temperature levels, when the rotational speed is adapted (approximately inversely proportional to the diameter) to obtain a constant pressure ratio. However, a large decrease in Reynolds number is unavoidable, resulting in higher viscous losses and a lower overall cycle efficiency. Due to these increased losses, the rim speed should even increase slightly with miniaturization in order to keep the pressure ratio constant. Figure 2 shows the cycle efficiency as a function of pressure ratio and turbine inlet temperature, calculated for compressor and turbine efficiencies of 80 per cent, corresponding to the latest designs. It is clear that the use of a recuperator is crucial for obtaining acceptable efficiency. A pressure ratio of 3 is chosen, a value still achievable with a single stage compressor. Below this value, efficiency drops sharply, whereas higher values offer smaller efficiency improvements. The mass flow is fixed at 20 g/s corresponding to the optimum specific speed of 0.7 [23]. It will require a three-dimensional impeller but allows for achieving a high compressor and turbine efficiency. Higher temperatures invariably boost the overall efficiency. In large turbines, the blades are cooled by internal cooling channels and protected by thermal

Calculated efficiency without (left) and with (right) recuperator (80 per cent efficiency for turbine and compressor, 70 per cent recuperator effectivity)


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barrier coatings. In the case of microturbines, internal cooling of blades a few millimetre in size is unrealistic. Thus, the turbine temperature must be lower, resulting in a lower cycle efficiency. However, higher allowable temperatures can be obtained by using ceramic materials as silicon nitride and silicon carbide, but this requires new manufacturing techniques. In the current design the turbine inlet temperature is set to 1200 K. This would give a cycle efficiency of 36 per cent, without taking into account bearing friction, thermal losses, pressure drops in the combustor and recuperator, and electrical losses. Another major consequence of the small dimensions is the extreme temperature gradient between the hot turbine and colder compressor. The resulting heat flux provokes a non-negligible decrease of both compressor and turbine efficiency [24, 25]. Heating the flow during compression decreases the efficiency because the mechanical energy input by the shaft is larger than that in the adiabatic case. An opposite effect occurs in the turbine where cooling of the flow during expansion decreases the energy that can be extracted. Another consequence of this large heat transfer is an increase of compressor exit temperature and a lower turbine exit temperature. The net effect is a decrease of the recuperator effectiveness. Because the microturbine presented in this paper is five times larger than that in MIT’s system [12], heat losses will have less effect on its efficiency [24]. Furthermore, a ceramic turbine impeller is used with a more than five times lower thermal conductivity (28 W/m/K) than the silicon used by MIT (149 W/m/K). A gasturbine’s net power output is the (small) difference between the large turbine power output and the large compressor energy requirement. The deterioration of aero-performance of the components with decreasing dimensions and increased heat transfer results in an even larger decrease of power output and cycle efficiency. A careful optimization is needed to guarantee positive output also at the smallest dimensions [26]. The most elaborated knowledge on small hightemperature turbomachines is in the field of turbochargers, operating at up to 200 000 r/min, and auxiliary power units in aeroplanes. All operations above this limit require new developments as well in terms of material and manufacturing techniques as in terms of aero-thermodynamics [27]. There is also no guarantee that the existing flow solvers and turbulence models will remain accurate for these extremely low Reynolds numbers. 4

compared to large-scale gasturbines to get a residence time in the combustion chamber that is high enough to allow a complete combustion of the injected fuel. As the flow velocity through the combustion chamber is of the same magnitude as in large gasturbines, although the length of the combustion chamber is much smaller, the residence time of the gas in the combustion chamber is reduced by almost an order of magnitude compared to the large counterparts. In this reduced time frame, air and fuel have to be mixed (at low Reynolds numbers) and combustion should be completed. Therefore, the size of the combustion chamber in microgasturbines is much larger than in large gasturbines in relative terms, and highly reactive fuels are used. For this project, hydrogen was chosen as a fuel for its fast mixing and combustion capabilities, besides its wide combustion limits. A second limitation to the design space of microcombustion chambers is the increased heat loss due to the higher surface-to-volume ratio and the higher thermal gradients compared to large gasturbines (in relative terms). This can result in flame quenching and will reduce the efficiency of the combustion chamber. Furthermore, the residence time constraint is aggravated by these higher losses as they lead to a reduced overall temperature in the combustor and thus also a lower chemical reaction rate. The inlet velocity of the chamber is a compromise between size and performance (pressure drop). For the microturbine under consideration, a value around 20 m/s is adopted resulting in an annular combustion chamber length of 50 mm and a height of 18 mm. As can be seen from Fig. 3, plates are added to the annular geometry to create extra recirculation zones and to increase the residence time of the gases inside the combustor. The first T-shaped plate is added for fuel injection and mixing whereas the second plate is added to stabilize and anchor the flame. Furthermore, a geometry with dilution is selected to increase the performance of the combustion chamber. A part of the air bypasses the flame

COMBUSTION

As already explained in section 2 of this paper, the relative size of the combustion chamber has to increase Proc. IMechE Vol. 221 Part C: J. Mechanical Engineering Science

Fig. 3

Combustion chamber layout JMES472 © IMechE 2007


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compressor and turbine, and from the rotor weight. The dynamic load originates from magnetic generator forces, rotor unbalance, and rotor vibrations. The stringent requirements of high rotational velocities, small dimensions, and high temperatures rule out the use of most conventional bearing types and turns the design of suitable bearings into a challenging task.

5.2

Fig. 4

Combustion simulation results (temperature scale in K)

and is injected downstream of this flame at both the external and internal diameter. In this way, both the stability and the efficiency of the chamber are increased. The bypass air cools the walls and in this way reduces the heat losses. Furthermore, as this (relatively) cold air is mixed with the combustion gases before entering the turbine, it allows the adoption of a higher equivalence ratio in the flame front while the turbine inlet temperature remains below its limit. Due to this higher equivalence ratio and the extra recirculation zones added by the injection of the bypass air, the stability of the flame is increased. Figure 4 shows a typical temperature profile obtained with Fluent for the current geometry. 5 5.1

BEARINGS Bearing requirements and operational conditions

One of the most remarkable effects of miniaturization is the rotational speed required by the thermodynamic cycle. This can be explained by the fact that the pressure ratio, which is independent from turbine size, is directly related to the circumferential speed. As a result, the rotational speed has to scale inversely proportional to the rotor diameter, resulting in speeds of more than 500 000 r/min for rotor diameters less than 20 mm. A reasonable overall efficiency can only be attained if the frictional losses in the bearings are kept to a minimum, despite the high velocity. The bearings must operate throughout the whole domain of possible temperature conditions. Steadystate temperatures between 100 and 1000 ◦ C can be expected depending on the exact location of the bearings. Continuous and stable operation at these elevated temperatures, as well as startup during transient temperature conditions must be ensured. The static bearing load mainly arises from pressure differences between inlet and outlet of both JMES472 © IMechE 2007

Overview of bearing types

The application of conventional plain or roller bearings is incompatible with almost every requirement stated above. Due to the high temperature condition, for example, it is practically impossible to provide lubrication for operation at high speeds throughout an acceptable lifespan. 1. Ball bearings exist for speeds up to 500 000 r/min (e.g. for dentist drills), and also for high temperatures (full ceramic bearings up to 800 ◦ C). However, the combination of high speed and high temperature is not feasible. 2. Magnetic bearings can be constructed using permanent magnets, electromagnets, or a combination of both. A drawback of permanent magnet bearings is the limited operating temperature due to demagnetization above the Curie temperature. Furthermore, it is not possible to build a stable magnetic bearing by using only magnets and ferromagnetic materials (Earnshaw’s theorem [28]). Electromagnetic bearings, on the other hand, are more complex and consume electrical energy, while the purpose of this microturbine generator is to generate electrical energy. 3. Air bearings seem most suitable for this application. They combine high speed and low friction and the temperature is only limited by the chosen bearing materials. Among air bearings there is a choice between aerostatic, aerodynamic, and hybrid bearings. Aerostatic bearings are fed with compressed air, which is in principle not available during startup of this autonomous power generator. Thus, aerostatic bearings have startup problems unless compressed air is stored in an additional reservoir. Once the turbine is rotating at sufficient speed, compressed air can be tapped from the compressor. Hydroinertia bearings, such as those developed by Isomura et al. [29] and Tanaka et al. [30], are a special type of aerostatic bearings. The bearing clearance is much larger than in normal aerostatic bearings. This results in a larger air consumption and supersonic flow, yielding a sub-ambient pressure and thus a suction force. With this bearing type, Isomura et al. [29] reached a maximum speed of 870 000 r/min for a mesoscopic turbine. Proc. IMechE Vol. 221 Part C: J. Mechanical Engineering Science

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Aerodynamic bearings generate a pressure profile by virtue of the relative shearing motion between two surfaces separated by a narrow non-parallel clearance. No external pressure supply is needed. One drawback of aerodynamic bearings is the phase of dry friction during startup. Another major drawback is the possibility of occurrence of self-excited instabilities that limit the maximum attainable speed. The key issue in applying aerodynamic bearings is assuring stable operation throughout the whole speed and temperature range. Most research groups MIT [12, 31], Onera [14], Tokyo [32], Belgium [33] active in the field of microgasturbines opt for aerodynamic bearings. Hybrid air bearings are a combination of aerodynamic and aerostatic bearings. During startup they act as aerodynamic bearings. At higher speed compressed air can be applied via the feed holes, which has a stabilizing effect at high rotational frequencies. Thus, the compressed air is only required when it is actually available. Air consumption is very low in comparison to the total compressor air flow, such that the impact on the global efficiency is extremely small. Stanev et al. [34] was able to operate a (large) spindle of 25.4 mm diameter up to 120 000 r/min or 3.0 × 106 DN (bearing diameter D in mm × shaft speed N in r/min). For the application presented here, aerodynamic bearings are the most promising choice on condition that the issue of instability is tackled.

5.3

Stability problems of aerodynamic bearings

Aerodynamic bearings are prone to self-excited instabilities, generally referred to as fractional speed whirl. This whirling of the rotor at approximately half the rotational speed limits the maximum attainable speed. Once this type of whirling has set in, it cannot be suppressed or overcome by further increasing the speed. In that respect it is not comparable to eigenfrequency phenomena. The onset speed of self-excited whirling can be predicted by looking at undamped eigenvalues of the motion equations [35]. The equations contain dynamic stiffness and damping coefficients of the bearings, which are normally calculated by a perturbation technique. The occurrence of self-excited whirling is related to the presence of cross-coupled stiffness and damping coefficients, as well as negative direct damping coefficients. It is generally known that an unloaded or lightly loaded aerodynamic bearing has a stronger tendency towards self-excited vibrations at high speeds [36]. A small bearing load results in a small eccentricity which in its turn results in large cross-coupling terms for stiffness and damping in comparison to the Proc. IMechE Vol. 221 Part C: J. Mechanical Engineering Science

direct terms. Because of the so-called ‘cube square law’, small-scale systems very often encounter this problem. The loading due to rotor mass scales down with L 3 and the bearing load capacity scales according to L 2 , with L as a characteristic length. 5.4

Stabilization methods

One of the most effective ways to improve rotor stability is to reduce the clearance to radius ratio due to its fifth power influence on stability [35]. However, the reduc tion of the clearance is limited by machining tolerances and the risk for crashing due to rotor vibrations or shock loads. It also leads to increased viscous losses. Other methods stabilize the rotor by eliminating or at least diminishing the cross-coupled terms with respect to the direct terms [31, 37]. 1. Increasing the load by applying an external stabilizing load in the form of pressure or magnetic attraction. This increases the eccentricity resulting in smaller cross-coupling terms in comparison to the direct terms. MIT tried to implement this method in their microturbine [31]. They aimed at eccentricity levels in the order of 0.9 and higher. With such high eccentricity the bearing is very vulnerable for crashes. 2. Introducing a small inflow of air (hybrid bearing). 3. Adding grooves on the bearing surfaces, or using multi-lobed geometries. A third way to stabilize the rotor is to add structural damping to the bearings or the supporting structure. The combination of these three stabilization methods (clearance, cross-coupling terms, and damping) can be realized by means of conformal bearings, especially flexible foil bearings [38]. This is explained in the next subsection. 5.5

Foil bearings

Foil bearings are known for their stability at high speed. Despite the numerous publications on foil bearings, these mainly focus on experimental work. To our knowledge, the stability of foil bearings has never been thoroughly analysed through numerical modelling. A few publications focus on partial analyses only [39]. As a result, hardly any design guidelines are available, making it difficult to scale down the current designs. It is often unclear why certain designs work well and what their operational limits are. Therefore, software tools were developed within the framework of this project to analyse the steady and dynamic behaviour of foil bearings. The model is based on the techniques mentioned in section 5.3 [33]. The major advancement brought by this tool is that it not only JMES472 © IMechE 2007


calculates bearing characteristics but also stability maps: critical mass plotted as a function of load and speed. These stability maps show that foil bearings do not have an ‘unlimited’ speed range, contrary to what is often claimed. The maximum speed is only increased by 50–100 per cent, compared to a rigid journal bearing with the same geometry [33]. This improvement can be attributed to the cross-coupling terms. However, foil bearings can safely operate with smaller clearances and higher eccentricities than rigid bearings without the risk of crashing: the foil gives way when the rotor touches or nearly touches the foil [33]. Eccentricities even larger than the clearance are used. Because the critical mass determining the stability has a fifth power dependence on the clearance, the stability and the maximally achievable speed can be considerably increased by using smaller clearances. Friction between different foil layers adds damping and thus further stabilizes the system. An additional advantage for this application is that foil bearings suffer less from rotor growth as a result of differential thermal expansion or centrifugal growth. 5.6

Future approach

Aerodynamic bearings with conformal surfaces present themselves as the most promising choice in achieving the stringent requirements. Current work focuses on designing new foil bearing concepts, suitable for the small dimensions encountered here. Hybrid bearings are chosen as a backup solution and for experimental purposes. Figure 5 shows the experimental set-up used to test the different bearing designs. It consists of a housing with nozzles and sensor interfaces in which different bearing modules can be inserted. In the picture shown, aerostatic thrust and journal bearing inserts are used. With this set-up a maximum speed of 425 000 r/min has already been attained.

Fig. 5

Modular air bearing setup (rotor diameter: 6 mm)


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ROTOR DYNAMICS

The microturbine should not operate close to any critical speed. In addition, any resonance that has to be passed during startup is a source of problems. Vibrations can be excited at a frequency synchronous with the rotational frequency, at a subsynchronous frequency, or at a supersynchronous frequency. When using air bearings, fractional speed whirling can occur resulting in subsynchronous shaft vibrations. The magnetic forces in the generator cause supersynchronous vibrations, with a frequency ratio depending on the number of generator poles, teeth, and construction. The magnetic forces do not only excite bending or suspension modes, but the ripple on the torque also excites torsional modes. Imbalance causes synchronous vibrations: below the resonance frequency, the rotor rotates around the bearing’s geometric centres; above the resonance frequency, the rotor rotates around its mass centre line (phenomenon of inversion). Asymmetries in the bearings, generator, and suspension as well as an eccentric position of the shaft may cause vibration at the synchronous frequency or a multiple of it, depending on the order of the asymmetry. In the ideal case, the critical speeds of the shaft should lie either at very low frequencies (a few 100 Hz) where they will do no harm due to the low speed, or at least a factor of 2, 3, 4, . . . above the rotational frequency, depending on the supersynchronous excitations. A stiff shaft suspended on air bearings has very low suspension eigenfrequencies and relatively high bending and torsion eigenfrequencies, exactly what is needed. However, the geometry of the shaft is not favourable for high resonance frequencies, as heavy masses are located at the ends and the compressor inlet introduces a low stiffness in the centre. In the current design the suspension eigenfrequencies are a few 100 Hz, whereas the first bending and torsion eigenfrequencies lie around 8.5 and 17.5 kHz, respectively. As the rotational frequency (8.3 kHz) lies very close to the first bending eigenfrequency, excitation of this mode will be a problem during steadystate operation. Current rotor optimization aims to increase the first bending eigenfrequency. During startup certain supersynchronous eigenfrequencies will inevitably be excited. The only solution is to pass these eigenfrequencies and to minimize their excitation, for instance by minimizing the generator ripple. To limit the rotor vibrations due to unbalance both at critical speeds and at operational (supercritical) speed, the rotor must be dynamically balanced in multiple planes: two for the suspension modes plus one for each bending mode below the operational frequency [40]. Due to the small bearing clearance, the difference between the bearing’s geometrical centre Proc. IMechE Vol. 221 Part C: J. Mechanical Engineering Science

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line and the rotor mass line should fall within the range of a few micrometre.

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COMPRESSOR AND TURBINE

As already mentioned, the efficiency of every component is critical for the overall gasturbine performance. This is particularly true for the compressor and turbine, requiring efficiencies of at least 60–70 per cent to guarantee a positive power output [41]. It is clearly a challenge to obtain this performance level because of the low Reynolds numbers, increased heat transfer, and lower relative geometrical accuracy of the small components. The latter leads to an increased gap between blades and shroud and hence more leakage flow and losses [25]. One way to reduce the impact of these losses on performance is by increasing the mass flow. The mass flow of the current design is therefore fixed at 20 g/s for an impeller diameter of 20 mm and a pressure ratio of 3. Two-dimensional impellers are easier to produce but have low specific speed. This means reduced performance and limited mass flow [23, 42]. Higher mass flows require a larger leading edge radius, which means shorter chord and larger blade height. The blades are subjected to increased shear and bending and a finite element stress analysis showed that this results in unacceptably large stresses in the blade root for the target pressure ratio. Increasing the blade thickness at the hub may alleviate the stress problem, however, at the expense of higher blade losses. The most effective way to reduce the root stresses is by using three-dimensional blades. This results in normal stresses and minimal bending because the centrifugal forces can be aligned along the blade height. A computerized design system, based on a genetic algorithm, artificial neural network, and threedimensional Navier–Stokes solver has been used to optimize the three-dimensional compressor and turbine geometry [43]. The 40◦ backward leaned impeller has seven blades and seven splitters. It is followed by a long vaneless diffuser with radius ratio 2 extending up to the recuperator inlet. The turbine nozzles create a 74.5◦ preswirl. The radial turbine impeller has eight blades and almost zero exit swirl at design point. A three-dimensional Navier–Stokes solver [44] predicts a total to total efficiency exceeding 80 per cent for both compressor and turbine. The heat transfer between the compressor and turbine is evaluated by a conjugate heat transfer calculation [20]. The impact on compressor and turbine efficiency is evaluated at less than 2 per cent points. In the current design of compressor and turbine, the centrifugal stresses easily rise to values of 1600 MPa for steel or nickel alloy rotors. The reason is the Proc. IMechE Vol. 221 Part C: J. Mechanical Engineering Science

high blade tip speed (524 m/s) in comparison to large turbines (300–400 m/s). This is a general trend in microturbines caused by the reduction in efficiency with the decreasing Reynolds number [45] and the relatively larger clearances between rotating and stationary parts (reduced manufacturing accuracy). For the compressor a special high-strength steel is chosen (Vascomax C-350) with a yield strength of 2.4 GPa at room temperature. Titanium alloys are an alternative as their lower density results in lower stresses. However, their low stiffness is a problem for the natural frequency of the complete rotor; the middle section (the compressor inlet) should be as stiff as possible. The material load is even more severe for the turbine as the material has to resist these high stresses at temperatures up to 1200 K. In large turbines, metal blades are used at much higher temperatures. This is possible thanks to the application of thermal barrier coatings, internal blade cooling, and air film cooling. In film cooling compressed air is injected along the blade surface, forming a cold boundary layer, thus separating the hot gas from the blades. However, these cooling techniques are not realistic for a rotor diameter of 20 mm. High-temperature high-strength alloys (e.g. Haynes R-41, Nimonic 115) quickly lose their strength above 1100 K. At 1200 K Haynes R-41 has a yield stress of 420 MPa, compared to 820 MPa at room temperature and 750 MPa at 1100 K. Ceramic materials are more temperature resistant and also have a lower density. The ceramic currently in use (Kersit 601, a Si3 N4 –TiN composite) has a density half that of the nickel alloys, thus the centrifugal stresses are about 800 MPa. As this corresponds to the material’s strength at high temperature, further optimization is required.

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RECUPERATOR

In order to reduce the overall fuel consumption, the effective use of energy is very important in the design of gasturbines. Heat recuperation is often used to improve the overall cycle efficiency of standard gasturbines. However, in small-sized gasturbines this improvement is much more questionable. Both achievable compressor pressure ratios and turbine inlet temperatures are significantly lower and pressure drops are much larger compared to conventionally sized gasturbines. The additional pressure drop introduced by the small channels in the recuperator should not undo the benefits of heat recuperation. Ideally, channels should be short and their number as large as possible within the constraint of a maximum allowable volume. Independent of gasturbine size, a thorough thermodynamic analysis [46] leads to a first recuperator JMES472 © IMechE 2007


design directive; pressure drop should be preferably located at the cold side of the recuperator. For this reason, the hot channels should be larger than the cold ones. The development of a microrecuperator with high effectiveness and low pressure drops is challenging, especially for small dimensions. In conventionally sized recuperators, complex, well designed fin configurations are used in order to improve the gas–air heat transfer. In order to avoid these costly and difficult to machine fin configurations, alternative recuperator designs are needed for microscale applications. Till now, no adequate microrecuperators have been tested in combination with a microgasturbine. MIT already proposed a microrecuperator design in 2001, but this was tested independently [47]. No experimental results are reported in literature on the improvement of the overall cycle efficiency because of the incorporation of a microrecuperator.

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GENERATOR

The electrical machine which is connected to the turbine, is designed for two objectives: as a startup motor for the turbine and as a generator when the turbine reaches steady-state regime. In contrast to normal designs with speeds of about 10 000 r/min and operation temperatures of up to 350 K, this machine will operate at speeds up to 500 000 r/min and elevated temperatures between 470 and 570 K due to the compactness [48]. Consequently, the position and the construction of the machine is an important issue. The high speed results in high mechanical stresses so that the geometry of the rotor is reduced to simple shapes. Another result of the high speed is the high operating frequency (a multiple of the rotational frequency). The high frequency introduces skin effect in the electrical circuit and eddy currents in the magnetic circuit. The high temperature results in the positioning of the permanent magnet away from the heat source. Permanent cooling is inevitable since permanent magnets lose their properties around 400 K (Curie temperature). Because of the higher resistivity of copper, the losses in the windings will be higher as well. It is obvious that both speed and temperature interact. In principle there is the choice between axial and radial machines (direction of the magnetic field lines in the gap). Axial machines generate high axial attraction forces between rotor and stator that have to be taken up by the bearings. For radial machines the magnetic forces counteract each other largely. Two radial machines are currently under investigation: a permanent magnet synchronous machine and a switched reluctance machine. The first one has permanent magnets in the rotor whereas the second JMES472 © IMechE 2007

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one has a solid steel rotor. Therefore, the switched reluctance machine has an inherently stronger construction towards centrifugal stresses. The magnetic flux through the machine can be created by coils actuated by an external energy source or by a permanent magnet. The use of coils may be interesting in high temperature environments, but causes additional energy consumption. The temperature problem of permanent magnets is bypassed by placing the magnet on the lowest temperature side of the turbine (the compressor side) and by passing the inlet air through cooling channels in the generator for additional cooling. 10

FABRICATION

The main challenge is the fabrication of threedimensional impellers for both compressor and turbine. Although lithographic techniques give the best results for precision and resolution, they are not able to produce the three-dimensional rotor geometries that yield better turbine performance. Therefore, lithography is only used for the smallest microturbines (4–10 mm) [12–15]. For the mesoscopic turbines (8–20 mm) other processes are used. Stanford [16] uses a gel-casting technique with a fugitive wax mould to produce a 12 mm diameter silicon nitride rotor. The mould is produced layer by layer by Shape Deposition Manufacturing. Tokyo University [17] and Tohoku University [18] use five-axis micromilling to produce metal impellers with a diameter of, respectively, 8 and 10 mm. The authors opted to use a metal compressor and a ceramic turbine. The compressor forms a monolithic metal part with the shaft and will be machined on a Kern five-axis micromilling machine. Figure 6 shows the result of a first milling test for the compressor, using milling tools down to 1 mm diameter. For this initial fabrication test, the geometry is simplified

Fig. 6

Five-axis milling test (diameter: 20 mm)

of

the

compressor


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by removing the smaller splitter blades. The final compressor will have splitter blades. The production of the turbine impeller is more difficult because a Si3 N4 –TiN ceramic composite (Kersit 601 from Saint-Gobain) is used for the reasons mentioned in section 7. One of the technologies under investigation for machining ceramic impeller prototypes is electro-discharge machining (EDM). Figure 7 shows a machining test of a single blade with threeaxis EDM milling. Copper milling electrodes of 3, 2, and 1 mm diameter are used. The work piece is clamped in a rotational indexing head to address the individual cavities. The only limitation to be machinable by EDM is that the material has to be sufficiently conductive. Kersit 601 can be easily machined by EDM due to the conductivity of the titanium nitride phase (0.0013 cm). However, this material faces two problems with respect to this application. The first is that at the high temperature of the plasma channel of the EDM sparks, silicon nitride decomposes and forms nitrogen bubbles. This results in a relatively high surface roughness, with a negative impact on flow losses. The second is that titanium nitride present in the composite oxidizes at high temperatures. The titanium present in the TiN diffuses towards the surface to form a layer of nearly pure TiO2 [49]. Oxygen diffuses inwards to form TiO2 and SiO2 . Important to notice are the voids formed in the affected layer and especially immediately under the TiO2 top layer.With time, these voids grow such that the TiO2 layer will break away. Therefore, Kersit 601 samples were coated with a 3 μm thick coating of pure silicon nitride by plasma-enhanced CVD. Figure 8 shows test results for uncoated and coated samples heated in a furnace at a temperature of 1100 ◦ C for 24 h in air. According to the results shown in Fig. 8 the uncoated sample is clearly affected by oxidation to a depth of 35 μm, the coated sample is clearly not affected.

Fig. 8

Uncoated (left) and coated (right) samples tested at 1100 ◦ C for 24 h in air

In the future, ceramic powder injection moulding is envisaged for large series production. Pure silicon nitride can then be used, which takes away the need for a coating.

11

CONCLUSION

Miniaturization of gasturbines is not straightforward and is not simply the scaling down of large gasturbines. The small dimensions have implications on the efficiency of the components due to the increased viscous friction. Also limits in fabrication accuracy cause additional flow losses. The allowable turbine inlet temperature is lower than for large turbines, making it more difficult to obtain positive cycle efficiencies. The impeller diameter of the gasturbine under development was set to 20 mm and the mass flow maximized towards 20 g/s to assure a positive cycle efficiency. A recuperator is added in order to improve cycle efficiency. For much smaller dimensions and mass flows the feasibility is less sure. The main mechanical problem is the required rotational speed of 500 000 r/min. Air bearings are the only feasible option regarding speed and temperature. Aerodynamic foil bearings and hybrid bearings have been selected. The challenge is then to assure stable operation despite the many possible disturbances caused by imbalance, shock load, self-excited whirl, and resonances.

ACKNOWLEDGEMENTS

Fig. 7 Three-axis EDM milling test for the turbine (diameter: 26 mm)


This research is sponsored by the IWT, the Institute for the Promotion of Innovation by Science and Technology in Flanders, Belgium, Project SBO 030288 and by the Belgian programme on Interuniversity Poles of Attraction (IAP5/06: AMS). Also, this work was carried out within the framework of the EC Network of Excellence ‘Multi-Material Micro Manufacture: Technologies and Applications (4M)’. Dr Jan Peirs wishes to acknowledge the Industrial Research Fund K.U.Leuven. JMES472 © IMechE 2007


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APPENDIX Notation CVD Dh DN-number EDM L RoDh = 2ωDh /W W

chemical vapour deposition hydraulic diameter bearing diameter in mm × shaft speed in r/min electro-discharge machining characteristic length rotational number relative flow speed

ω

rotational speed
