A LOW PRESSURE RATIO MICRO GAS TURBINE FOR DOMESTIC COMBINED HEAT AND POWER

Alister Clay*, G.D Tansley

Mechanical Engineering & Design, Aston University, Birmingham, UK *claya@aston.ac.uk

Abstract: This paper presents work on an alternative approach in delivering a 1 kW micro gas turbine (MGT) for domestic combined heat and power (DCHP). A method of reducing stress in a centrifugal compressor impeller is presented followed by prototype manufacture for test by direct metal laser sintering (DMLS). Low pressure ratio is assessed on a system perspective with regard to a novel, coiled pipe-in-pipe recuperator. The pressure ratio must be sufficiently high if better performance is to be achieved. Using a lower specific speed can significantly reduce shaft speed if a higher dynamic pressure component is accepted. Keywords: micro, gas, turbine, centrifugal, impeller, stress, CFD, FEA, recuperator, DMLS MOTIVATION

The motivation for a low pressure ratio system is described herein: Direct: Higher component efficiency [1] Broader operating range [2]. Lower speed: Low risk commercial ‘off-the-shelf’ bearing solution. Improved non contact bearing stability [3] Larger size: Reduced aerodynamic losses, less aggressive loading. Reduce tip clearance losses. Better manufacturing tolerances. Low specific speed: Use kinetic energy from low diffusion ratio to stabilize flow. INTRODUCTION

Although favorable in terms of life and performance, air bearings remain unproven commercially at these scales. ‘Off-the-shelf’ oil cooled journal bearings found in turbochargers are a lower risk commercially but with a maximum shaft speed of approximately 220,000 rev/min [4] pressure ratio and so engine performance are limited.

From an impeller geometry algorithm written in Matlab a simple compressor impeller CFD optimization study [5] found two impellers with opposing characteristics: impeller a 170,000 rev/min 15° backsweep and impeller d 200,000 rev/min 54° backsweep but similar performance: pressure ratios between 2.1 – 2.2, isentropic efficiencies between 71% - 72%. Both impellers were characterized by low diffusion ratios: 1.13 and 0.88 and low specific speed: 0.41 and 0.47 below the 0.6 – 0.8 range proposed by [6].

The CFD study assessed the impact of blade backsweep and rotational speed to produce a pressure ratio of 2.15 with a fixed impeller diameter of 40mm for retrofit on to the Garrett GT1241 bearing platform. Lower rotational speeds were achieved by reducing blade backsweep and blade height. An off-design study demonstrated a wide operating range and regular flow, both features thought to be the result of a low diffusion ratio and pressure ratio. STRESS Problem

To accommodate stress during the CFD study, the

discharge tangential velocity component, 2U , was maintained below 470 m/s to permit the use of AlC355 T-6 alloy [7] a common cast material for centrifugal compressor impellers found on turbochargers. To retro fit the new impellers on the existing bearing platform the back face profile and bore diameter were measured from the GT1241 compressor impeller and modeled on the new impellers a and d.

The stress condition was assessed using a 1° hub segment of impeller d. FEA conducted at 200,000 rev/min found a stress concentration around the bore edge on the back face. Although design similitude was carried over, the von Misses stress was greater than the material strength, suggesting failure. This prompted an investigation into the geometric condition of the impeller. Material

Before investigating the back face geometry, a study into the most common compressor materials [7] was carried out, see Table 1. The maximum stress exceeded both the yield and tensile stress regardless of the material used.

Whilst bearing tests up to 220,000 rev/min were confirmed, it’s possible the smallest production turbochargers rarely exceeded rotational speeds beyond 160,000 rev/min during application. Further FEA conducted in this investigation suggested maximum speed to be around 150,000 rev/min using AlC356 T-6 material. Table 1: Maximum von Misses stress results at bore edge on back face. Material (17-4PH) (AlC356

T-6) (AlC2618 T-61)

(Ti 6Al-4V)

ρ(kg/m3) 7833 2680 2760 4429 σt(N/m2) 1.207e+09 2.280e+08 4.410e+08 8.274e+08 σy(N/m2) 1.137e+09 1.850e+08 3.720e+08 1.050e+09 σMax 2.100e+09 8.836e+08 9.100e+08 1.379e+09 Back face geometry

With a fatigue limit, 17-4PH stainless steel was

selected as the new material. The prospect of a continuous running operating strategy and numerous start stop cycles, suggested aluminum alloys without a fatigue limit would unsuitable. Cast metals with higher melting temperatures such as stainless steel are also favored due to less shrinkage. Certain geometric restrictions were imposed during the back face geometric study to ensure the impeller would sit in the machined profile on the journal bearing platform:

• A flat radial surface to allow contact with the

thrust bearing. • A tangential arc from the impeller tip to the flat

surface of the thrust bearing face. • 0.644 mm rotor tip width, approximately the same

as the GT1241 compressor impeller. Material was introduced to form a slope from the

bore edge to the thrust bearing face, see Fig. 1, to re-distribute the back face mass. The stress concentration around the bore edge was reduced by 55%. The stress is shown to be more evenly distributed across through the hub core and back face. The same triangulated profile was removed from the thrust bearing by spark erosion at HCM Engineering Ltd1 to maintain the axial distance between the bearing impeller.

Fig. 1: Centrifugal stress condition of Impeller d with triangulated thrust bearing profile. Blade root

A 30° (360°/blade number) impeller segment was modeled to assess the stress condition of the blade root, see Fig. 2. Cyclic symmetry was maintained as blade position was mirrored on each segment face. To maintain the aerodynamic composition of the impeller channel, geometric alterations were kept to a minimum. For the splitters, constant radius fillets were introduced at the root. For the blades, a variable radius fillet was fitted wherein material along the blade root increased with radial distance. This method maintained the sharp leading edge whilst providing adequate stress distribution towards impeller exit where the centrifugal loading is largest. There were no requirements to add rake or reduce the blade number.

1 HCM Engineering Ltd, Pedmore Road, Lye, Stourbridge, West Midlands, DY9 7DZ. Telephone : +44 (0) 01384 422 643 | Fax: +44 (0) 01384899210 | Email: info@hcmeng.co.uk

With subtle geometric alterations maximum stress was within a material yield safety factor of 2. The centrifugal loadings combined an angular velocity of 200,000 rev/min with a radial acceleration from 0 to 200,000 rev/min in 1 sec.

Maximum resultant displacement was 0.036 mm using stainless steel at 200,000 rev/min, well within the 0.3 mm tip clearance used for the CFD study.

Fig. 2: Equal blade segment showing Von Misses stress distribution with 100 times deformation. MANUFACTURE Impellers

The impellers were manufactured by Direct Metal Laser Sintering (DMLS) at Materials Solutions2 . Materials Solutions specializes in metal powder bed Additive Layer Manufacturing (ALM) using a fibre laser equipped EOS M270 machine. ALM manipulates CAD files to build 3D objects in 20 micron layers. Metal powder is melted with the laser which builds the part in successive layers to achieve a 100% core density. After the DMLS process the part undergoes a heat treatment process prior machining followed by grit blasting. Final parts have metallurgical properties similar to forgings and surface finish similar to castings. Various metals are available including stainless steel 17-4, however these prototypes were manufactured with a high temperature nickel superalloy C263 more common in gas turbine combustors. It has a density of 8200 kg/m3, tensile strength of 950 MPa and a yield strength 0.2% PS of 675 MPa. The benefits of DMLS include: • Efficient manufacture of working prototypes from

CAD files similar to rapid prototyping. • Delicate three dimensional thin wall manufacture

ideal for small turbomachinery components. • Overhung features.

2 Materials Solutions, Unit 8, Great Western Business Park, McKenzie Way, Worcester, WR4 9PT. Telephone: +44 (0) 1905 732160 | Fax: +44 (0) 1905 530224 | Email: info@materialssolutions.co.uk

0.03086 mm

0.01016 mm

0.02280 mm

0.01987 mm

• A 25 impeller batch DMLS time would be 72 hours. Prototype process time was 8 hours + 2 days heat treatment + 5 minutes grit blasting + 5 minute machine time on the back face.

Obstacles overcome during manufacture include: • Tapered as opposed filleted inlet blade edges. • Inclusion of internal thread within bore. • Skewed/overhung blades

After finding a method of preventing damage to

the thin inlet blade tips from the DMLS powder re-coater, Fig. 3 (left), Material Solutions successfully built the impeller geometry. A grit blasting method to remove loosely adhered particles without causing damage, Fig. 3 (right), to the thin blades was also developed. Essentially, the micro compressors were built without compromise and accurately replicated the original aerodynamic and stress design intent. The final impellers are shown in Fig. 4.

Fig. 3: Manufacturing challenges overcome.

Fig. 4: Final micro impeller pieces. Shrouds

Each shroud was manufactured on a Fanuc Robodrill alpha-T21iE 3 axis CNC milling machine. An overhang volute was built as per the restrictions of a 3 axis machine. Fig. 5 shows a sectioned view of the bearing assembly, a drinks can is added for reference.

Fig. 5: Cross section of bearing assembly showing shroud and profile of the thrust bearing. SYSTEM

A simple coiled pipe-in-pipe recuperator micro gas turbine system has been assessed as a potential MGT recuperator solution using a pressure ratio of 2.15, see Fig. 6. The recuperator seeks to utilize standard stock materials and manufacturing techniques to provide a simple, low risk, solution at an expected cost of 1.5 times material cost [8].

At five coil turns the recuperator is a sufficient size to house and contain the other MGT components within the central void as per the design intent and produces a MGT thermal efficiency of 11.3%, below the target value of 15%. At one coil turn, the MGT thermal efficiency is 8%, around half the target and at twenty coil turns the MGT thermal efficiency is 16%. Using an optimized pipe diameter ratio the recuperator performance is maximized within the overall volume envelope permitted leaving component efficiency and or pressure ratio to raise MGT thermal efficiency.

Fig. 6: Indicative exploded view of MGT

In MGTs, the optimum pressure ratio for

maximum thermal efficiency is much lower compared to conventional gas turbines with higher component efficiencies and smaller relative pressured drops [9]. The thermal/hydraulic performance of the 5 turns recuperator was 47.63% effective and 6.33% total system pressure drop. Based on system calculations using the above recuperator performance, the optimum pressure ratio is approximately 4.0 which yielded the target thermal efficiency of 15%. The target thermal/hydraulic performance was 75% effective and 10% total system pressure drop which would yield the target thermal efficiency of 15% using a sub optimum pressure ratio of 2.15.

Fig. 7 shows calculated values for impeller d if it were designed to deliver a pressure ratio of 4.0. Due to a low specific speed the shaft speed is still only 280,000 rev/min which although is beyond the realms of oil cooled journal bearings, it easily achievable with air or electromagnetic bearings. From a DCHP perspective with requirements for low maintenance, non contact bearings would be welcomed provided there is a commercial advantage or design necessity.

Low specific speed impellers are thought to compromise efficiency [10]. With the low diffusion

characteristic which accompanies low specific speed, raising pressure ratio will produce an ever increasing dynamic pressure component as suggested by Fig. 8 which may also compromise combustion downstream.

To accurately assess the effect of pressure ratio on recuperation, CFD simulations were performed on the coiled pipe in pipe recuperator with 5 turns at increasing pressure ratios. The boundary conditions used included mass flows from a system analysis and static pressure values from compressor mean line analysis of impeller d [5], see Fig. 7. Fig. 9 shows the effect of the different pressure ratios in terms of recuperator and system performance.

Whilst the dynamic pressure component increases proportionately with pressure ratio, to maintain the same power output the increased pressure ratio is met with a reduction in mass flow on a systems level, which causes an overall reduction in flow speed on entry to the combustor. Due to a large hydraulic diameter in the recuperator, the fluid speed altogether is relatively slow which is also beneficial for heat exchanger effectiveness.

150

200

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2.0 2.5 3.0 3.5 4.0Pressure ratio

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Fig. 7: Calculated values for a 40 mm diameter, 54° backswept compressor impeller at increasing pressure ratios.

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/m2)

Total 220 Total 200 Total 170 Static 220 Static 200 Static 170

Small dynamic pressure components from lower pressure

ratios

Lower speed, lower backsweep = larger dynamic pressure

components

Pressure component/Shaft speed Fig. 8: Pressure components of low specific speed impellers with different backsweep and shaft speed.

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System efficiencyRecuperator mass flowCombustion inlet velocity

Fig. 9: Affect on MGT system performance using a coiled pipe in pipe recuperator of at 5 turns. CONCLUSION

Low pressure ratio is assessed on a system perspective with regard to a novel, coiled pipe-in-pipe recuperator. The pressure ratio must be sufficiently high if better performance is to be achieved. Using a lower specific speed can significantly reduce shaft speed if a higher dynamic pressure component is accepted. Using a low pressure ratio won’t eliminate difficulties but shifts challenges to other areas. REFERENCES [1] Japikse, D., Centrifugal Compressor Design and

Performance. 1996, Vermont: Concepts ETI, Inc. [2] Came, P.M. and C.J. Robinson, Centrifugal

Compressor Design. IMECHE, 1999. 213: p. 139-155.

[3] Peirs, J., et al. Experimental verification of compressor performance for an ultra-micro gasturbine. PowerMEMS. 2009. Wachington DC, USA: Imperial College London.

[4] Garrett. Garret Product Catalog, Turbochargers, GT12 Family, GT1241-756068-1. 2005.

[5] Clay, A. and G.D. Tansley, A micro gas turbine for UK domestic combined heat and power. Journal of Power and Energy - Part A, 2010. 224(6): p. 839-849.

[6] Rodgers, C. Specific speed and efficiency of centrifugal impellers. Proceedings of the 25th IGTC. 1980. New Orleans, USA: ASME

[7] Japikse, D., Introduction to Turbomachinery. 2nd ed. 1997, Vermont & Oxford: Concepts ETI. Inc & Oxford Univerisity Press.

[8] McDonald, C.F., Low-cost compact primary surface recuperator concept for microturbines. ATE, 2000. 20(5): p. 471-497.

[9] Clay, A. and G.D. Tansley, An Analysis of Micro Gas Turbines for UK Domestic Combined Heat and Power. IDGTE, Power Engineer, 2010. 14(3): p. 18-34.

[10] Casey, M.V., Computational methods for preliminary design and geometry defintion in turbomachinery AGARD paper LS-195. Turbomachinery design using CFD, 1994. May: p. 1-22.