NUMERICAL SIMULATION OF UNSTEADY- FLOW PROCESSES IN WAVE ROTORS. Florin Iancu

Transcription

1 Proceedings of IMECE ASME International Mechanical Engineering Congress November 13 19, 2004, Anaheim, California USA IMECE NUMERICAL SIMULATION OF UNSTEADY- FLOW PROCESSES IN WAVE ROTORS M. Frackowiak Warsaw University of Technology Institute of Aeronautics and Applied Mechanics 24 Nowowiejska Str. Warsaw, Poland Pezhman Akbari Michigan State University Dept. of Mechanical Engineering 2500 Engineering Building East Lansing, Michigan Phone: +1 (517) Fax: +1 (517) Florin Iancu Michigan State University Dept. of Mechanical Engineering 2500 Engineering Building East Lansing, Michigan Phone: +1 (517) Fax: +1 (517) Norbert Müller Michigan State University Dept. of Mechanical Engineering 2455 Engineering Building East Lansing, Michigan Phone: +1 (517) Fax: +1 (347) A. Potrzebowski Warsaw University of Technology Institute of Aeronautics and Applied Mechanics 24 Nowowiejska Str. Warsaw, Poland Janusz Piechna Warsaw University of Technology Institute of Aeronautics and Applied Mechanics 24 Nowowiejska Str. Warsaw, Poland Phone: +48 (22) Fax: +48 (22) ABSTRACT The wave rotor (pressure exchanger) is a device working based on a relatively simple idea of operation, but is challenging in its technical realization and difficult to simulate numerically. It has been common practice to create and use specialized codes for simulating the wave rotor operation. The current work presents an attempt of developing 2D and 3D models of radial and axial wave rotors using the commercial software package FLUENT. In this study geometrical models are used for the device casing and rotor cells. The application of carefully chosen initial and boundary conditions enabled the realization of relative motion of the rotor model. The vast information about the unsteady processes occurring during simulation are visualized. It occurs that such type of models are useful for the final test of devices, after the geometry was optimized by the use of specialized but much simpler 1D codes. Keywords: wave rotor, wave disc, radial wave rotor, shock waves, CFD INTRODUCTION Wave rotor technology has shown significant potential for increasing efficiency and performance of thermodynamic systems. The wave rotor is an unsteady-flow device that utilizes shock waves to exchange energy from a high energy fluid to a low energy fluid, increasing both temperature and pressure of the low energy fluid. The major principle underlying the operation of wave rotors is based on the physical fact that when two fluids with different thermodynamic properties are brought into direct contact for a very short time, pressure equalizations occurs faster than mixing [1]. This novel technology has been used in various applications that include the use as a supercharging device for IC engines [2-28], a topping component for gas turbines [29-34], and in refrigeration cycles [35-38], and more [39]. Wave rotors have been a research goal for decades, inspired by Burghard patent in 1929 [40]. However, difficulties mainly related to poor knowledge about unsteady-flow processes limited the dissemination of wave rotor concept until World War II when Seippel in Switzerland [41-44] implemented this concept into a locomotive gas turbine [45]. Since then, numerous research efforts have been carried out to overcome the challenges prevented the commercialized applications of wave rotors devices. Most of these efforts were 1 Copyright 2004 by ASME

2 experimental and thus expensive. Computational methods and digital computer facilities were too little developed in the past, so extensive theoretical methods required to improve the progress were found too difficult. Recent advances and experiences obtained by the wave rotor community have renewed interest in this technology. These advances include new computational capabilities allowing accurate simulation of the flow field inside the wave rotor, and modern experimental measurements and diagnostic techniques. Several precise numerical codes have been developed, allowing accurate simulation of flow field inside the wave rotor. However, most of these codes are not commercially available, hence a few groups of researchers who have developed these codes are using them. A broader accessibility of appropriate CFD tools could facilitate a wide range of wave rotor analysis The development of modern multi-purpose commercial software packages has reached a level that allows the successful modeling and analysis of the operation of many technical devices, including wave rotors. Several commercial codes are now available that can be applied to investigate many problems related to wave rotor design and operation. Such codes are particularly interesting for 2D and 3D modeling of full devices and some special problems. They offer tools that allow for relatively easy geometric preparation, a range of typical boundary conditions, relatively fast and robust solving, and a wide range of post-processing which is valuable for engineers and scientists. Yet, the use of such codes is not as fast as the use of common office software. Although the geometry and boundary conditions can be modeled relatively fast, using such complex commercial software packages, the computational effort is still enormous, so that flow field is often only available after lengthy, time-consuming computations. Therefore, such simulations are not suitable for an initial geometry search or a geometry optimization but can be performed as a last stage of investigation, verifying solutions of particular problems or the full operation of a complete wave rotor. For preliminary investigations, initial design, and optimization they are not necessarily as efficient as specialized codes. CURRENT AND PAST WAVE ROTOR FLOW SIMULATIONS Spalding at Imperial College is one of the pioneers using a 1D method to investigate unsteady flows inside wave rotors in the early 1970s. He formulated a numerical procedure for wave rotors considering the effects of heat transfer and friction. It utilized novel features to ensure solutions free from instabilities and physical improbabilities, as reported by Azoury [36]. Based on this numerical model, a computer program was developed by Jonsson [46], and it was successfully applied to pressure exchangers [47-49]. In 1981, Turbopropulsion Laboratory (TPL) at Naval Postgraduate School (NPS), directed by Shreeve, started an extensive numerical and analytical wave rotor program to evaluate the wave rotor concept and its potential application in propulsion systems [50]. For numerical simulations, two different approaches to the solution of the unsteady Euler equations were examined in the overall program. First, Eidelman developed a 2D code based on the Godunov Method to analyze the flow in wave rotor channels [51-54]. Unlike contemporary one-dimensional approaches [55], the 2D code showed the effect of gradual opening of the channels. The main conclusion of these studies is that if the channels are straight, the flow remains nearly 1Dl, which in turn leads to minimal mixing losses caused by rotational flow in the channels [56]. However, when the channel of the wave rotor is curved, even an instantaneous opening of the channel does not lead to the development of a 1D flow pattern with small losses. Computation time of such a 2D code has been reported to be quite long. Therefore, interest was given to development of a 1D code introduced by Mathur based on the Random Choice Method solving the Euler equations [57, 58]. The code which is called WRCOMP (wave rotor component) is a first order accurate in time and it was unconditionally stable. WRCOMP calculated the unsteady process inside the wave rotor, inlet and outlet opening times and other useful design parameters required for a preliminary design. The outputs from WRCOMP are used in a second program, called ENGINE, performing a performance calculation for a turbofan jet engine [59-61]. The results have also confirmed the significant performance improvement that could be expected integrating a wave rotor into a turbofan engine. Work was planned toward incorporating the effects of friction and heat transfer into WRCOMP and also including other engine configurations in ENGINE code. Some modifications to WRCOMP code was later started [62, 63], but further development was not pursued after terminating the wave rotor research around During 1990s, a few numerical investigations of pressure wave superchargers are reported. Nour Eldin and his associates in Germany have developed a fast and accurate numerical method for predicting unsteady flow field in pressure wave machines, using the theory of characteristics [64-70]. Piechna et al. in Poland have developed experimentally validated 1D and 2D numerical codes to analyze the flow field inside the Comprex [71-77]. Piechna has also proposed a compilation of the pressure exchanger with internal combustion wave rotor presenting the idea of the autonomous pressure wave compressor [78]. By initiating a wave rotor research at NASA Lewis Research Center (now Glenn Research Center) in the late 1980s, Paxson developed a quasi-one-dimensional gasdynamic model to calculate design geometry and off-design wave rotor performance [79-80]. The code uses an explicit, second order, Lax-Wendroff type TVD scheme based on the method of Roe to solve the unsteady flow field in an axial passage for timevarying inlet and outlet port conditions. It employs simplified models to account for losses due to gradual passage opening and closing, viscous and heat transfer effects, leakage, flow incidence mismatch, and non-uniform port flow field mixing. Recent improvement and validations have completed it as a preliminary and general design tool [82-89]. Welch has also established 1D and 2D analysis models to estimate the performance enhancements of wave rotors [90-92]. Larosiliere has also established a multi-dimensional code to investigate the effect of gradual opening and losses [93-95]. 2 Copyright 2004 by ASME

3 Motivated by NASA wave rotor successes, Lear at the University of Florida initiated analytical and numerical methods to investigate different configurations of wave rotors. His team developed an unsteady two-dimensional numerical code using a direct boundary value method for the Euler equations to analyze the flow in wave rotors and their adjoining ducts, treating the straight or curved channel walls as constraints imposed via a body force term [96]. The code was later used to simulate the flow filed of the three-port NASA [97-99] wave rotor A parametric study of gradual opening effects on wave rotor compression processes is reported, too [100]. Fatsis and Ribaud at the French National Aerospace Research Establishment (ONERA) have also developed a 1D numerical code based on an approximate Rieman solver taking into account viscous, thermal, and leakage losses [101, 102]. The code has been applied to three-port, through-flow, and reverse-flow configurations. Nagashima et al. have developed 1D [103] and 2D [104] CFD codes to simulate the flow fields inside through-flow four port wave rotors, including the effects of passage-to-passage leakage. The codes have been validated with experimental data obtained by a single-channel wave rotor experiment WAVE ROTOR SIMULATIONS WITH COMMERCIAL CFD PACKAGES Some attempts at simulating pressure wave superchargers with the help of commercial codes already have been undertaken. One such code is GT-POWER, in which pipe elements have been divided into a series of objects for which the conservation equations have been solved. By dividing variables into primary (density, total internal energy) and secondary variables (pressure, temperature) a staggered grid has been the result of discretization. Sets of elementary elements can be connected into one net, controlling the flow between them. The pipe elements can also include wall friction and heat transfer. Proper modules can represent local changes of pipe cross-section or valving. An interesting description of techniques for optimization of timing, shaping, and control of pressure wave changers using GT-POWER has been described by Podhorelsky et al. [105]. For the preliminary tests results generated here the commonly available CFD software FLUENT has been used. The present work shows 2D results of simulating complete operation of a novel radial wave rotor (wave disc) with straight channels and curved channels. Details of the operation of an aerodynamic speed control and the oblique opening of curved channels are described as well. Furthermore, 3D results of a complete conventional axial wave rotor are presented proving the capability of this software. For most of the results investigated here, no experimental data exist for possible verification. The only concept for which data are available is the conventional axial wave rotor. It seems that such type of models are useful for the final tests of devices for which the geometry has been already optimized by the use of specialized, but much simpler, 1D codes. The presented results shall not be interpreted as a proposition of a particular geometry for Figure 1: Wave disc with straight channels practical applications. While they can be a base for further investigations, they rather present some illustrations of physical phenomena generated in different configurations and phases of operation and showing the capability of the code. WAVE DISC WITH STRAIGHT CHANNELS For wave discs, centrifugal acceleration acting on the fluid in the channels and increasing with radius needs to be considered in governing equations additionally. In a proper port configuration the presence of centrifugal forces can improve the scavenging process which is a typical challenge in wave rotor design and operation. Increase of wave disc rotational speed increases centrifugal forces and reduces the number of ports which can be located in a disc of same diameter. Hence, for higher speed, instead of two port sets and two cycles per revolution only one port set and one cycle per revolution could be used. The model considered here is a reverse-flow wave disc with 60 straight channels as shown in Fig. 1. The disc has an inner radius of 0.15 m and an outer radius of 0.3 m, rotating clockwise at 8330 rpm. Two sets of four ports, as in a conventional four-port pressure exchanger, are used. It is assumed that the fresh air enters the channels with a temperature of 300 K at a pressure of 10 5 Pa. It is compressed by a high-pressure, high-temperature gas entering at 1000 K and Pa. The gas expands and leaves at the lower value 10 5 Pa. The geometry of the ports is not optimized. Due to relatively low rotational speed of the wave disc and the position of inlets and outlets, the optimal inclination of the port walls are different than in common axial-flow configurations. The presented results serve only for verification of the general idea of wave disc operation where our knowledge is still very limited. Figure 2 shows some preliminary results. Relatively uniform regions shown before and after the four ports indicate 3 Copyright 2004 by ASME

4 a) b) c) d) e) f) g) h) Figure 2: Contour plots, velocity vectors, and particle paths (time=8.6000e-04s) a) velocity magnitude (m/s); b) velocity vectors colored by velocity magnitude (m/s); c) radial velocity (m/s); d) tangential velocity in (m/s); e) static pressure (Pa); f) static temperature in K; g) and h) path lines colored by particle ID 4 Copyright 2004 by ASME

5 that the combination of diameters, speed, and port arrangement is not optimized, because nearly no changes are seen in these regions. Only the static pressure contours in Fig. 2-e show a certain radial stratification in these regions indicating the action of centrifugal forces. Figure 2-e also clearly shows the effects of compression and expansion waves and how the low-pressure region is created by supporting the ingestion of fresh air. Figures 2-b and 2-f show the flow casing especially for the burned gas which is a typical feature of reverse flow configurations. Fig 2-g shows recirculation in the middle of a channel almost closed at both ends. Furthermore, Fig. 2-a and 2-c show effects of gap leakage, especially on the left side before the high-pressure gas port. WAVE DISC WITH CURVED CHANNELS Reduction of the wave disc diameter can be obtained by the use of curved or spiral channels. Curving the channel allows for modulating the acceleration in flow direction that results from the acting centrifugal forces, since only the component of the centrifugal force that is tangential to the path of the fluid motion affects the fluid acceleration. Thus the channel shape, which may be easier to modify in a wave disc configuration, is another parameter that should be taken into account for a proper wave disc development. Figure 3 is the configuration used in the numerical simulations presented here. The disc radii and number of channels are the same as for the wave disc with straight channels, as explained before. The rotational speed now is 4000 rpm. While the temperature boundary conditions for the ports also are the same as before, now two different high-pressure levels are used. In the lowerright corner Pa is set for both high-pressure gas inlet and high-pressure air outlet ports while Pa for these ports in the upper-left corner. Figure 4 shows similar plots as in Fig. 2, but now for Pa pressure. Regimes of the curved channel configuration are shown and complimented with plots of the radial velocity component at the inner and outer disc radii. The static pressure contour in Fig. 4-d shows a pressure peak even above the inlet pressure, short before opening of the highpressure air outlet port. The static temperature contour in Fig. 4-e shows a similar casing as Fig. 2-f, but with deeper gas penetration and an unexpected carry on of expanding hot gas stretches after the exhaust gas port opens. At the same time fresh and colder air bypasses these strikes and reaches apparently to the exhaust port as confirmed by comparison with the velocity vectors in Fig. 4-b and the distribution of the radial velocity in Fig. 4-c. The latter actually reveals some inwards recirculation of hot gas that is supposed to move outwards to the gas outlet. The same effect can be seen in the radial velocity diagram for the outer diameter of the disc (Fig. 4-g) right after the peak of outwards flow with about 250 m/s. The same plot shows an analogous pattern for the regime with Pa at the high-pressure ports. Figure 4-f reveals similar recirculation on the inner disc diameter which is the air side. Both plots in Fig. 4-f and 4-g show clearly the 2D feature of the channel flow with much lower flow speeds at the channel wall boundaries. Using only a few channels in the model reduces calculation time and allows simple tests for the compression and expansion process with a much finer grid. This method is used to find a rotational speed at which the compression and expansion process is better tuned with the port opening and closing fixed in this study. This way a rotational speed of 8300 rpm was found appropriate and the results are represented in Fig. 5 and 5. The static pressure distribution in Fig. 5-e shows the traveling of a primary chock wave in the upper two channels and the shock reflection in the third channel below. The distribution of the local temperature in Fig. 5-f shows now clearly the complicated shape of the interface between hot gases and cold air. Again a deep penetration of hot gases can be seen. This relates not only to the high-pressure ratio of 3:1, but also due to the huge incidence angle between the incoming hot gas and the channel wall direction. It generates a very low pressure at suction side (trailing side) of the channel ends and in turn a recirculation with high-speed inwards flow of hot gas and outwards sucked air are created. Figures 5a, 5-d, and 5-g give further evidence of this phenomena, indicating that the application of 1D models would result in rather error here. These strong recirculation and resulting considerable nonuniformity can probably minimized by using a similar backward swept channels for a through-flow configuration where high-pressure gas and fresh air are introduced at the inner diameter and high-pressure air and expanded gas are scavenged through outer ports. Otherwise the flow entering from the outer diameter would need to have a considerable huge tangential flow component in rotational direction to match the now relative high speed (8300 rpm) of the disc. Figure 3: Wave disc with curved channels 5 Copyright 2004 by ASME

6 a) b) c) d) e) f) g) Figure 4: Low and high-pressure parts of wave disc with curved channels: contour plots and velocity vectors (time=2.817e-3s): a) velocity magnitude (m/s); b) velocity vectors colored by velocity magnitude (m/s); c) radial velocity (m/s); d) static pressure (Pa); e) static temperature in K; f) radial velocity at inner diameter; g) radial velocity at outer diameter. 6 Copyright 2004 by ASME

7 a) b) c) d) e) f) g) Figure 5: High-pressure part of wave disc with curved channels: contour plots and velocity vectors (time=8e-4s): a) velocity magnitude (m/s); b) velocity vectors colored by velocity magnitude (m/s); c) radial velocity (m/s); d) tangential velocity in m/s; e) static pressure (Pa); f) static temperature in K; g) Zoom in high-pressure gas inlet of (b) 7 Copyright 2004 by ASME

8 Using the same geometry as above, but now for an outwards flow that results from different boundary and initial conditions, the low-pressure process is investigated in which the expanded gas is scavenged and fresh air is ingested. Figure 6 shows much less non-uniformities in the channels than Fig. 5. Figures 6-a through 6-d show actually very well matching flow between the rotating channels and the ports. This is especially the case for the upper channels two and three and perfect matching between the air inflow and the channel wall between the lower two channels. This is reflected even in distribution of the static temperatures in Fig. 6-f, where only a typical 2D effect is seen at the air inlet that results from the ingested air jets and some wall heat transfer. The static pressure distribution in Fig. 6-e shows a homogenous distribution in the various phases of expansion, also reflected in the decreasing temperature shown in Fig. 6-f. The calculation of the complete model with curved channels like described further above was repeated with rotational speed of 8300 rpm. Again the lower high-pressure ports are Pa and the upper are Pa, both parts being independent. The results are presented in Fig. 7, showing that compression and expansion are generally working. The static pressure contours in Fig. 7-d shows again a radial pressure stratification in the regions where both ends of the channels are closed as discussed for Fig. 2-e. With higher rotational speed now the pressure difference is even higher and the pressure plot in Fig. 7-g shows a pressure increase at the outer diameter of about Pa and simultaneously a pressure decrease of about 10 4 Pa at the inner diameter in the regions x/l and 0.8 1, mainly due to the action of centrifugal forces. This way the fresh air is pre-compressed before the high-pressure gas enters. The complete temperature contours in Fig. 7 now show a very deep penetration of hot gas due to the phenomena explained before for Fig. 4 and 5. However, the penetration is obviously deeper at the right side where the high-pressure level is Pa. While centrifugal acceleration acts on both sides the same way, the compressed air at the right side has much greater density at Pa and is hence propelled much stronger along the trailing wall (suction side) of the channel, which creates a stronger recirculation almost along the full channel length. This is subsidized very clearly by Fig. 7b and 7-c which show respectively the outwards directed and inwards directed radial component of the flow. This indicates again very strongly the problem of inwards scavenging of compressed air here in combination with a strong incidence angle for the inflow of the high-pressure gas. Despite that, Fig. 7-b and 7-f indicate outflow for the full length of the gas outlet port, the port could be much shorter if only gas is to be scavenged and channel flushing with fresh air is not desired. The temperature contours in Fig. 7-e indicate this clearly with the end of the outflow of hot gas. More fresh air follows from the fresh air inlet port due to the action of centrifugal forces, like in a backwards swept turbo-compressor. The near zero value of the radial velocity component through the outer diameter at x/l 0.2 and 0.7 in Fig. 7-f could mark the beginning of the fresh air flushing. AXIAL-FLOW WAVE ROTOR Some 3D calculations were performed for an axial reverseflow wave rotor suitable for a microturbine. Its preliminary design was widely obtained by an analytical procedure by Akbari and Müller [106]. Results are shown in Fig. 8. While Fig. 8-a and 8-c show clearly again the hot gas casing typical for reverse-flow wave rotors, Fig. 8a and 8-b indicate very nicely the effects of compression and expansion waves. Keeping the same speed and diameter the rotor could accommodate two cycles per revolution. If ducting allows reverse arrangement of the cycles might be desirable for more homogeneous cooling of the rotor. OBLIQUE CELL OPENING While some 1D models can be constructed for typical straight channels that open gradually and perpendicular to the channel center line, considerable 2D effects are expected when curved channels are used, since their opening and closing is oblique to the channel center line. For this reason the opening of the gas outlet port was investigated for curved channels on the above wave disc. Velocity contour plots are shown in Fig. 9, sorted by channel position in respect to the leading edge of the outlet port. Besides clear 2D effects these contour plots also show leakage flow in the gap between the disc and the casing. JET ACCELERATION PASSAGES Port timing in conventional wave rotor configurations is already difficult to optimize. While the shock wave front is sharp and its position can be predicted relatively easy, the expansion wave is more a fan and predicting its boundaries is more challenging. Furthermore, both types of waves can be reflected by different types of boundaries and they can interact with each other. Simple wave rotor configurations without compensating pockets can generally work properly only in very narrow range of rotational speeds. Jet acceleration passages for auto-tuning of rotational speed are investigated as an example of a particular problem. For a conventional axial wave rotor, a model with up to five neighboring channels, high-pressure gas inlet, and compressed air outlet port was considered as shown in Fig. 10. Several test calculations with different rotor velocities are performed and aerodynamic forces were recorded. Since initial gas parameters were held constant, with different rotor speeds, the shock wave generated by opening the channel to the high-pressure gas port arrives at the opposite channel end at different times meaning at different location in respect to the stationary end plate. Figure 11 shows results of the acceleration passage operation for three different tangential rotor velocities of 30, 40, and 50 m/s. The slowest velocity corresponds to the case in which the shock wave activates operation of the accelerating passage. The velocities in the passage are then relatively high. Similar modeling can be carried out for such a passage that acts in the opposite direction as a jet break, decelerating the rotor speed if it is too fast. 8 Copyright 2004 by ASME

9 a) b) c) d) e) f) Figure 6: Low-pressure part of wave disc with curved channels: contour plots and velocity vectors (time=8e-4s) a) velocity magnitude (m/s); b) velocity vectors colored by velocity magnitude (m/s); c) like b) but fresh air inlet only; d) like b) expanded gas outlet only; e) static pressure (Pa); f) static temperature in K; 9 Copyright 2004 by ASME

10 a) b) c) d) e) f) g) h) Figure 7: Low and high-pressure part of wave disc with curved channels: contour plots and velocity vectors (time=1.732e-3s): a) velocity (m/s); b) radial velocities outwards (m/s); c) radial velocities inwards (m/s); d) static pressure (Pa); e) static temperature in K; f-g) for inner and outer diameter versus non-dimensional circumferential length starting at the opening of lower right high-pressure gas port increasing in rotational (clock-wise) direction: f) radial velocity component g) static pressure; h) static temperature. 10 Copyright 2004 by ASME

11 a) b) c) d) e) f) g) Figure 8: Axial wave rotor: contour plots and velocity vectors (time=2.05e-3s) a-c) Static Pressure (Pa); d) Velocity Vectors colored by velocity magnitude (m/s); e-g) (time=2.5524e-3s):e) local density (kg/m 3 ); f) static pressure (Pa); g) static Temperature in K 11 Copyright 2004 by ASME

12 c) j) g) d) a) h) e) b) i) f) Figure 9: Velocity contour plots (m/s) for oblique channel opening ordered by channel position: a) time 5.09e-4s; b) time 5.89e-4s; c) time 6.69e-4s; d) time 7.49e-4s; e) time 8.29e-4s; f) time 9.09e-4s; g) time 9.89e-4s; h) time 1.069e-3s; i) time 1.149e-3s; j) time 1.229e-3s 12 Copyright 2004 by ASME

13 Figure 10: Jet acceleration passage: velocity contours (m/s) (time=4e-4s) CONCLUSION Commercial CFD packages offer tools that allow for relatively easy preparation of geometries, some range of typical boundary conditions, relatively fast and robust solvers and a wide range of post-processing which are valuable tools for engineers beside scientists. Using the commonly available software package FLUENT it has been shown extensively that such tools can be effectively used for simulation of rather complex time dependent flow phenomena that occur typically in wave rotors. It appears that the novel concept of wave discs and other particular problems can be investigated by 2D models, whereas the conventional axial-wave rotor should be treated as a 3D model. While geometry and boundary conditions can be modeled relatively fast, using such complex commercial software packages the computational effort is still enormous, so that flow field is often only available after hours. Therefore, such simulations are not suitable for an initial geometry search or a geometry optimization but can be performed as a last stage of investigation, verifying solutions of particular problems or the full operation of a complete wave rotor. For preliminary investigations, initial design, and optimization they are not necessarily as efficient as specialized codes. REFERENCES [1] Gyarmathy, G., 1983, How Does the Comprex Pressure-Wave Supercharger Work?, SAE Paper [2] Barry, F. W., 1950, Introduction to the Comprex, Journal of Applied Mechanics, pp [3] Berchtold, M., Gardiner, F. J., 1958, The Comprex: A New Concept of Diesel Supercharging, ASME Paper 58-GTP- 16. [4] Lutz, T. W., Scholz, R., 1968, Supercharging Vehicle Engines by the Comprex System, The Institution of Mechanical Engineers. [5] Zehnder, G., 1971, Calculating Gas Flow in Pressure- Wave Machines, Brown Boveri Review, No. 4/5/71, 4019 E., pp [6] Berchtold, 1974, The Comprex Pressure Exchanger: A New Device for Thermodynamic Cycles, JSAE Paper, Tokyo. [7] Doerfler, P. K., 1975, Comprex Supercharging of Vehicle Diesel Engines, SAE Paper Figure 11: Jet acceleration passage: velocity vectors colored by magnitude (time and legend in Fig. 10 applies) [8] Groenewold, G. M., Welliver, D. R., and Kamo, R., 1977, Performance and Sociability of Comprex Supercharged Diesel Engine, ASME paper 77-DGP-4. [9] Schwarzbauer, G. E., 1978, Turbocharging of Tractor Engines with Exhaust Gas Turbochargers and the BBC- Comprex, The Institution of Mechanical Engineers, Paper C69/78, pp [10] Kollbrunner, T. A., 1980, Comprex Supercharging for Passenger Diesel Car Engines, SAE Paper [11] Jenny, E., Zumstein, B., 1982, Pressure Wave Supercharger of Passenger Car Diesel Engines, The Institution of Mechanical Engineers, Paper C44/82. [12] Walzer, P., Rottenkolber, P., 1982, Supercharging of Passenger Car Diesel Engines, The Institution of Mechanical Engineers, Paper C117/82, London. 13 Copyright 2004 by ASME

14 [13] Zehnder, G., Mayer, A., 1984, Comprex Pressure- Wave Supercharging for Automotive Diesels - State-of-the- Art, SAE Paper [14] Berchtold, M., 1985, The Comprex, Proceeding ONR/NAVAIR Wave Rotor Research and Technology Workshop, Report NPS , pp , Naval Postgraduate School, Monterey, CA. [15] Spinnler, F., Jaussi, F. A., 1986, The Fully Self- Regulated Pressure Wave Supercharger Comprex for Passenger Car Diesel Engines, The Institution of Mechanical Engineers, Paper C124/86. [16] Jenny, E., Naguib, M., 1987, Development of the Comprex Pressure-Wave Supercharger: In the Tradition of Thermal Turbomachinery, Brown Boveri Review, 74, No. 8, pp [17] Mayer, A., 1988, The Free Running Comprex - A New Concept for Pressure Wave Supercharger, SAE Document PC 55. 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