## Introduction

Microturbines—sub-MW gas turbines—are a competitive solution to other power units, especially in mobile applications, because of their high power density, reliability, and low maintenance requirements. Gas turbines are also fuel flexible which makes them particularly suitable for future applications requiring lower emissions levels. On the downside, microturbines in a simple Brayton cycle configuration suffer important losses when increasing the operating pressure ratio, counteracting other possible advantages [1].

When operated in a recuperated Brayton cycle, microturbines have the potential to reach 40% thermal efficiency while keeping their other advantages. To do so, microturbines must operate at a low pressure ratio (between 5:1 and 10:1) and at a turbine inlet temperature (TIT) around 1300 °C and above [2,3].

In the last decades, the TIT of large gas turbines (>1 MW) has dramatically increased through innovations in cooling technologies, which also drove turbine component prices to high levels. These technologies are hard to implement cost-effectively in microturbines.

In an attempt to solve both efficiency and cost challenges of microturbines, tremendous development efforts have been made to develop monolithic ceramic gas turbines. Since the 1970s, monolithic ceramic properties, cost as well as manufacturing processes have significantly improved thanks to many substantial development programs. With these improvements in material quality and reliability, some programs successfully tested engines above 40% thermal efficiency, but never reached the required reliability and foreign/domestic object damage resistance to become products [46].

The failure of these massive research programs indicates the need for a drastic design change for monolithic ceramic turbine to succeed. An interesting and largely unexplored solution to solving the issues of monolithic ceramic turbine is the use of an inside-out configuration.

The inside-out architecture was first introduced by Kochendörfer et al. [7] in the 1970s with an all-ceramic rotor surrounded by a carbon fiber rotating structural shroud. The use of a ceramic hub and a ceramic cooling ring led to poor reliability, but still demonstrated the interest in this novel architecture.

More recently, inspired by the R4E engine [8], an ICT rotor comprising a superalloy flexible hub and cooling ring was developed and tested [9]. The prototype was tested at high temperature with no significant failure. However, tip speed limitations and crack propagations at the bottom of the blade were noted by the authors as challenges to address. Two major issues emerged from the design of a radially flexible hub for an ICT: (i) high tensile stresses in the hub and (ii) a high pressure interface at the bottom of the ceramic blade which results in local tensile stresses in the blade.

To address these issues, a sliding-blade turbine concept was proposed by Landry et al. [10]. The first step of this architecture exploration was the validation of the rotordynamics of the architecture, which was carried out successfully. It was concluded that rotordynamics stability in this configuration requires contact between the blade and the sliding plane in all conditions. However, the concept was not designed for nor tested under high-temperature turbine operating conditions.

This paper presents (1) the design of a sliding-blade ICT wheel for high temperature operation, (2) the experimental validation at a TIT of 1100 °C and a tip speed of 350 m/s, and (3) the ceramic blades life expectancy analysis. Experimental results and life predictions show that, with incremental improvements, the sliding-blade inside-out ceramic turbine concept is a viable architecture to significantly increase the TIT over state-of-the-art microturbines, providing gains in both efficiencies and power density.

The sliding-blade architecture eliminates the source of both hub and ceramic blade tensile stresses that were problematic in previous ICT designs by using a rigid hub with inclined planes, which serve as tracks for the blades to slide up and down. Contact between the hub and blade sliding planes is achieved with the axial force applied by springs to the blades. The friction between the blade and the structural shroud, combined with the force applied by the springs, maintains the integrity of the turbine in all conditions. The prototype design is showed in Fig. 1.

Fig. 1
Fig. 1
Close modal

Blade motion is dictated by expansion of the structural shroud, due to centrifugal loading. As rpm increases, the shroud radius increases, and the blades go up their respective sliding planes, as shown in Fig. 2. The sliding-blade architecture uses an axial spring comprised of a star-shaped disk with one spoke per blade. This ensures that each blade receives independent force and maintains contact with the hub when moving along its sliding plane, which is key to maintaining the structural integrity of the turbine wheel.

Fig. 2
Fig. 2
Close modal

## Prototype

In order to test the proposed design, a functional prototype was built, as shown in Fig. 3. The prototype is made of:

Fig. 3
Fig. 3
Close modal
• A structural shroud made from filament wound carbon fiber reinforced polyimide with 1800 MPa ultimate hoop strength and a 335 °C glass transition temperature made by Proof Research Advanced Composites Division (Dayton, OH).

• Axial springs, a cooling ring, and a hub made from Inconel 718.

• Monolithic ceramic blades made from Kyocera (Somerset, NJ) SN235P silicon nitride.

A small turbine outside diameter (75 mm) is used to lower development cost, but scaling up is possible up to a couple of MW turbines. This sliding-blade ICT prototype is the first of its kind to combine both a functional design and turbine grade materials that are able to withstand a TIT of 1100 °C at a tip speed of 350 m/s.

## Experimental Setup

The validation of the performance of the ICT prototype was carried out in a laboratory turbine environment. The ball bearings and compressor wheel are taken out of a Garrett GTX2860R automotive turbocharger unit and retrofit into a custom water-cooled housing. The can combustor is fueled with propane and a hydrogen pilot flame. An array of four K-type thermocouples is placed at the end of the combustion chamber to measure TIT. A remote-actuated bleed valve positioned between the compressor and the combustion chamber is used to control mass-flow. Cooling is provided by an external compressor. The cutaway computer aided design view and a photo of the test rig are showed in Figs. 4 and 5.

Fig. 4
Fig. 4
Close modal
Fig. 5
Fig. 5
Close modal

## Methodology

A two-phase test schedule was carried out to validate the prototype. First, a series of five 5-min tests were run, gradually increasing TIT from 950 to 1100 °C and tip speed from 300 to 350 m/s. Second, an hour-long continuous test was run at 1100 °C TIT and 350 m/s tip speed. During each test, TIT was ramped up to the test temperature in 15 s at a maximum tip speed of 120 m/s (40 krpm). This state was kept constant for 40 s before ramping up to test target speed in 60 s. The test was timed from that point on. At test completion, temperature and rpm were reduced to avoid hot soaking before cutting fuel off. For each test, an average cooling mass flow of 11% of the principal flow was maintained in the cooling ring. This cooling mass flow was used for proof-of-concept stage, but can be largely reduced in future turbines.

Reported TIT are averaged across the four thermocouples placed at the end of the combustion chamber.

Thermochromic paint from Thermal Paint Services is used to measure the temperature of critical turbine parts. Thermochromic paint KN5 is chosen to observe a wide range of temperature (260–1250 °C). Paint color after testing is compared to calibrated coupons. Each coupon is cured as per manufacturer recommendation for high temperature operation (1 h at 260 °C) and then brought up to calibration temperature for 5 min. Calibration coupons from the manufacturer are in accordance with in-house testing and are showed in Fig. 6.

Fig. 6
Fig. 6
Close modal

Thermal paint is airbrushed onto all ceramic and metallic components of the turbine wheel.

## Results

All tests ran smoothly and no problems or damage occurred. The rpm and TIT plots of the 1 h continuous test are shown in Figs. 7 and 8.

Fig. 7
Fig. 7
Close modal
Fig. 8
Fig. 8
Close modal

After the 1-h continuous test, visual inspection and disassembly of the prototype revealed no damage. A tested blade is shown in Fig. 9.

Fig. 9
Fig. 9
Close modal

This achievement represents a major step forward in the development of the ICT configuration. Tests are limited in inlet temperature because of the maximum operating temperature of the test rig, but there were no indications of blade or turbine degradation.

Figure 10 shows the temperature observed on the core components of the turbine wheel after testing at 1100 °C for 1 h.

Fig. 10
Fig. 10
Close modal

No coloration change in the paint occurred in the carbon fiber shroud, meaning this critical component is sufficiently cooled and remains under the curing temperature of the paint (260 °C), therefore well under its glass transition point (335 °C). The cooling ring also remains relatively cold with a maximum temperature of 600 °C at the interface with the blade. Blade temperature ranged 600–900 °C.

## Ceramic Blade Life Expectancy Analysis

In order to assess the life expectancy of the tested prototype, thermomechanical finite element (FE) simulations of the turbine are performed.

An FE model is used to analyze temperature distribution and identify critical stress locations in the turbine wheel and determine blade reliability. ansys Workbench is used to carry out a coupled thermal-structural FE analysis. The thermal setup is shown in Fig. 11. TIT used for the analysis is 1100 °C. Surface averaged coefficients of convection applied to the blade profile (840 $W/m2 °C$) and blade shroud (490 $W/m2 °C$) are extracted from Ref. [11] while coefficient of convection for the cooling ring (1250 $W/m2 °C)$ are extracted from a one-dimensional heat transfer model based on theory in Ref. [12]. Since the cooling ring is three-dimensional printed, the effect of high surface roughness due to this manufacturing process is also considered in the calculation of the convection coefficient based on the work of Stimpson et al. [13]. Inlet temperature of the cooling ring is 77 °C, increasing to 150 °C at outlet temperature. In a real engine with a pressure ratio around 8, the cooling ring air supply might need to go through an intercooler. A fixed temperature of 227 °C is applied at the base of the hub and the shaft based on oil temperature measurement made at bearing location. All other surfaces are considered adiabatic. The resultant temperature distribution across the turbine is shown in Fig. 12 and is consistent with temperature measurements obtained with the thermal paint.

Fig. 11
Fig. 11
Close modal
Fig. 12
Fig. 12
Close modal

Thermal analysis results are imported into the structural analysis. Frictionless contacts are used at all metal-to-metal interfaces to ease convergence of the model. A rough contact with radial press fit of 75 μm is used between the cooling ring and the structural shroud. The initial press fit and the high pressure at this interface prevents the cooling ring from slipping relative to the structural shroud. A coefficient of friction of 0.3 is used at the blade (Si3N4) to cooling ring (Inconel 718) interface, based on the work of Sliney and DellaCorte [14]. The resulting maximum principal stresses are shown in Figs. 13 and 14.

Fig. 13
Fig. 13
Close modal
Fig. 14
Fig. 14
Close modal
Creep issues resulting from these compression stresses are negligible at this temperature for a 5000 h operation time [15]. The reliability of the blades for slow crack growth is analyzed with cares software. Initially developed by NASA, cares is an integrated package that predicts the probability of failure of a ceramic component with respect to imported stress levels calculated by commercial fea software. The probability of failure Pf was computed according to the principle of independent action (PIA) theory. This theory stipulates that Pf for a given element is the product of Pf caused by each principal stress on the element. Component-level Pf is then obtained from the product of all elemental Pf. The equation used to compute Pf of a component is given by Nemeth et al. [16]
$pf(tf)=1−e1σ0V∫V∑iψidV$
(1)
in which
$ψi=(σiNtfB+σiN−2)mN−2$
(2)

for i = 1, 2, 3.

Static fatigue parameters presented in the work of Choi et al. [17] where used. Material constants estimate where provided by Kyocera. The resulting probabilities of a single blade failure (Pf) for different operating times are illustrated in Table 1.

Table 1

Blade probability of failure and target values

1 h100 h1000 h5000 h
Target Pf (estimated range)10−7 to 10−810−5 to 10−610−4 to 10−55 × 10−4 to 5 × 10−5
16 blade core Pf (excluding top 1 mm)2 × 10−91 × 10−77 × 10−73 × 10−6
16 blade set Pf3 × 10−51 × 10−31 × 10−26 × 10−2
1 h100 h1000 h5000 h
Target Pf (estimated range)10−7 to 10−810−5 to 10−610−4 to 10−55 × 10−4 to 5 × 10−5
16 blade core Pf (excluding top 1 mm)2 × 10−91 × 10−77 × 10−73 × 10−6
16 blade set Pf3 × 10−51 × 10−31 × 10−26 × 10−2

The tensile stress profile in the blade as seen in Fig. 13 clearly shows that stresses are localized at the high-pressure interface between the Inconel cooling ring and the ceramic blade. This is mainly the result of the difference in coefficients of thermal expansion of the two materials. This problem has been documented and solved in the past in a gas turbine application by reducing friction [19]. It is thus plausible to reduce friction and stresses at this interface. The same FE analysis was run at different friction coefficients (μ = 0.1–0.3) to observe its effect on blade Pf after 5000 h of operation. The results are shown in Fig. 15.

Fig. 15
Fig. 15
Close modal

The trend in numerical results for blade failure in relation to the friction coefficient of the cooling ring to ceramic interface reveals that a small reduction in friction at the interface leads to an exponential drop in failure probability. This indicates that the reduction of friction at this interface could eventually lead to a blade failure probability acceptable for a real engine application.

## Conclusion and Perspectives

A sub-MW turbine with typical gas turbine advantages and thermal efficiency over 40% has long been a target for the power market. The ICT configuration is a promising solution to this problem. The introduction of the sliding-blade architecture solved the most important issues encountered by previous ICTs.

This new ICT architecture has led to a prototype that successfully underwent a continuous test of more than 1 h at a TIT of 1100 °C and a tip speed of 350 m/s (115 krpm) without any failure or damage to the turbine.

These achievements mark an important milestone in the development of the ICT, but it is clear that tensile stress at the high-pressure interface at the top of the ceramic blade has to be reduced to achieve acceptable reliability. A numerical sensitivity analysis revealed that the reduction of friction at this interface would lead to an acceptable blade probability of failure. Overcoming this challenge could lead to an engine that opens new possibilities for applications requiring the power density, efficiency and reliability of large gas turbines, but at a smaller scale.

Further design iterations and longer testing of the sliding-blade architecture will be carried out to explore the performance limits of the ICT configuration in terms of tip speed and inlet temperature.

## Acknowledgment

The authors would like to express their regards for the technical support provided by the research professionals and professors of Université de Sherbrooke, École de Technologie Supérieure—LAMSI and Exonetik Turbo involved in the development of the ICT.

## Funding Data

• This work was funded by Exonetik Turbo, the Natural Science and Engineering Research Council of Canada (NSERC) CRD PJ 477320 14 (Funder ID: 10.13039/501100000038), and Defence Research and Development Canada (DRDC) Contract W7714 196710/001/SV (Funder ID: 10.13039/501100002956). Scholarships were provided by the Fonds de recherche du Québec–Nature et technologies (FRQNT).

## Nomenclature

• ICT =

inside-out ceramic turbine

•
• m =

Weibull modulus

•
• N, B =

static fatigue parameters

•
• Pf =

probability of failure

•
• tf =

time at which failure probability is evaluated

•
• TIT =

turbine inlet temperature, °C

•
• σ0V =

material specific constant

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