Research Papers: Gas Turbines: Structures and Dynamics

Predicting Gas Leakage in the Rotary Engine—Part II: Side Seals and Summary

[+] Author and Article Information
Mathieu Picard

Department of Mechanical Engineering,
Massachusetts Institute of Technology,
77 Massachusetts Avenue,
Cambridge, MA 02139
e-mails: mpicard@mit.edu;

Tian Tian

Department of Mechanical Engineering,
Massachusetts Institute of Technology,
77 Massachusetts Avenue,
Cambridge, MA 02139
e-mail: tiantian@mit.edu

Takayuki Nishino

Mazda Motor Corporation,
3-1 Shinchi, Fuchu-cho, Aki-gun,
Hiroshima 730-8670, Japan
e-mail: nishino.tak@mazda.co.jp

1Present address: Assistant Professor, Department of Mechanical Engineering, Université de Sherbrooke, Sherbrooke, QC J1N 0T2, Canada.

Contributed by the Structures and Dynamics Committee of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received September 10, 2015; final manuscript received October 2, 2015; published online November 17, 2015. Editor: David Wisler.

J. Eng. Gas Turbines Power 138(6), 062504 (Nov 17, 2015) (8 pages) Paper No: GTP-15-1445; doi: 10.1115/1.4031874 History: Received September 10, 2015; Revised October 02, 2015

The Wankel rotary engine offers a greater power density than piston engines, but higher fuel consumption and hydrocarbon emissions, in large part due to poor gas sealing. This paper presents a model for the deformable dynamics of the side seal, which completes a set of modeling tools for the comprehensive assessment of the gas leakage mechanisms in the rotary engine. It is shown that the main leakage mechanisms for the side seals are: (1) opening of the inner flank due to the contact with the trailing corner seal, (2) flow through the gap with the leading corner seal, (3) simultaneous opening of both inner and outer flanks due to body force at high speed, and (4) running face leakage due to nonconformability at high speed. The leakage mechanisms are qualitatively validated at low speed with observed oil patterns on the rotor from laser-induced fluorescence (LIF) experiments. Finally, the predicted total leakage area for all the gas seals ranges from 1.5 mm2/chamber at low speeds to 2 mm2/chamber at high speeds, which is in agreement with the previous experimental studies, and the three gas seal types (side seals, apex seals, and corner seals) each accounts for about 1/3 of the total leakage, with minor variation as a function of speed.

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Fig. 1

The rotary engine has a longer sealing line than the conventional reciprocating piston engines

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Fig. 2

The model finds the displacements of each cross section starting from a curved beam as reference position: (a) reference position and (b) calculated displacements of each cross section

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Fig. 3

The displacement of every cross section is defined by three translations (radial, axial, and circumferential) and one rotation (tilt): (a) reference position, (b) radial displacement (yg), axial displacement (zg), and tilt (αg), and (c) circumferential displacement (xg), length of the seal (L), and coordinate along the seal (s)

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Fig. 4

Circumferential forces are added on the side seals to calculate the circumferential displacements: (a) circumferential friction force on running face, (b) circumferential seal–groove friction force on seal flanks (shown for clockwise motion of the seal relative to the groove), and (c) asperity contact normal and friction forces on seal ends

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Fig. 5

Running face wear is approximated as parabolic

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Fig. 6

The model calculates the clearance at both ends of the seal as a function of the initial clearance and the displacements

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Fig. 7

Gas flow through the flank, running face, and end clearances is taken into account in the model: (a) flank and running face gas flow shown on a section view and (b) end leakage on a side view

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Fig. 8

As shown by the results of relative flank leakage at 2000 rpm full-load, predicted leakage can vary significantly as a function of discretization

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Fig. 9

Predictions at 2000 rpm full-load show that: (a) groove pressure follows chamber pressure, (b) the seal is pushed toward the trailing corner seal, (c) at high pressure, center of the seal (s/L = 0.5) is pushed toward the inner flank, but the trailing end (s/L = 1) remains in contact with the outer flank, (d) tilt remains small due to the restoring moment of the curved beam, and (e) flank and leading end leakage follow chamber pressure

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Fig. 10

Groove flank leakage is the most important leakage mechanisms at 2000 rpm full-load followed by seal end leakage

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Fig. 11

At low gas pressure (a), the seal does not fully conform to the distorted shape of the side housing. At high gas pressure (b), the gas force is sufficient to close most of the running face clearance. (a) 90 CA and (b) 390 CA.

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Fig. 12

Friction with the side housing pushes the seal toward the trailing corner seal. The reaction force with the trailing corner seal opens the inner flank at the trailing end. The counterclockwise displacement opens the gap at the leading end.

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Fig. 13

The predicted shape of the face seal at high pressure (390 CA) shows the opening of the inner flank at the trailing end

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Fig. 14

Predictions at 8000 rpm full-load show that: (a) pressure lag remains small, (b) the seal switches from one corner seal to the other due to circumferential body force overcoming friction during part of the cycle, (c) body force changes the radial displacement significantly, (d) tilt is increased, but the outer groove remains open, and (e) flank leakage is increased at the beginning of compression, running face leakage becomes important, and trailing end gap of the seal also leaks

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Fig. 15

At high speed, running face equivalent leakage area becomes important and flank leakage area is increased (running face leakage is calculated by averaging the results for the four housings, which have different calculated distorted shape)

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Fig. 16

At high speed, the seal does not fully conform to the distorted shape of the side housing, even with a high groove pressure: (a) 90 CA and (b) 390 CA

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Fig. 17

At the beginning of compression (240 CA shown here), the trailing half of the seal is pushed toward the outer flank by body force, which causes the two flanks to open simultaneously: (a) seal deformation and (b) force per unit length distribution along the seal

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Fig. 18

The two dominant side seal leakage mechanisms at low speed can be seen by observing the oil pattern on the side of the rotor by an LIF method (grayscale shows the oil film thickness, the brightest being the thickest oil film)

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Fig. 19

At low speeds (2000 rpm shown here), predicted leakage is almost equal between apex, corner, and side seals

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Fig. 20

At high speeds (8000 rpm shown here), apex seal flank leakage and side seal running face leakage become important. Side seal equivalent leakage area is also increased.



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