Regenerated Turboshaft Engines for Ground Power Plants Topped with Four-Port Wave Rotors

Regenerated Turboshaft Engines for Ground Power Plants Topped with Four-Port Wave Rotors

Fatsis Antonios

Mechanical Engineering Department Technological Educational Institute of Sterea Ellada

34400 Psachna, Greece

AbstractThe reduction of fuel consumption in gas turbines, along with the achievement of high levels of performance, remains one of the main goals of the engines designer. Addition of a heat exchanger (regenerator) is a common method adopted primarily in ground power units where engines weight and size is not an issue. Moreover, wave rotor technology has shown very promising results in terms of performance enhancement when applied to gas turbines. In this article, detailed calculations are performed for one and two-shaft regenerated gas turbine engines topped with wave rotors built for industrial applications. Having identified the proper placement of the wave rotor in such engines, thermodynamic cycle analysis is implemented, illustrating performance benefits in terms of specific work and specific fuel consumption,

Keywords Wave Rotor, Gas Turbine, Turboshaft, Ground Power Unit, Regenerator, Specific Work, Specific Fuel Consumption.

1. INTRODUCTION

   A wave rotor device is composed of a rotating and a stationary part. The rotating part consists of a number of axial straight blades between two coaxial cylinders forming a drum. The drum rotates inside a stationary casing having two end walls equipped with ports or manifolds, allowing only partial – due to openings on the end walls – inflow and outflow through the blade channels, as illustrated in Fig. 1. Through rotation, the channels extremities are periodically exposed to the ports located on the upstream and downstream stationary walls initiating compression and expansion waves which are propagated inside the channels. Thus, there is an unsteady pressure exchange between air and exhaust gas by means of pressure waves, as mentioned in [1]. The rotor has axial blade channels, thus assuring port flow alignment.

   Fig.1. Four-port wave rotor schematic configuration

   As a result the power required to keep the rotor at the designed speed is negligible, as mentioned in [2]. It only needs to overcome rotor windage and friction losses. It has been proven in [3] that for

   the same inlet and outlet Mach numbers the pressure gain in time- dependent flow devices can be much greater than in the ones operating in steady flow.

   When a wave rotor is integrated in a turbomachine, extra air compression is accomplished by means of compression waves formed in the wave rotor channels when hot exhaust gases coming out of the combustion chamber come in contact with air from the compressor. These compression waves are propagated inside the rotor channels in one direction, creating simultaneously expansion waves moving in the opposite direction. Consequently, hot gases are expanded inside the rotor and directed towards the turbine for further expansion. Even though propagation of compression and expansion waves creates a macroscopically unsteady flow pattern inside the wave rotor, the flows entering and exiting the rotor in the ports of the stationary manifolds are almost steady, apart from the fluctuations due to the gradual opening and closing of the rotor channels as they enter and exit the ports. A wave rotor succeeds both compression and expansion within a single component, contrary to the operation of a steady-flow turbomachine.

   The configuration assuming unchanged baseline compressor and turbine inlet temperature provides almost the best gas turbine performance enhancement of the wave rotor topped engine. However, the combustor could work under both higher pressure and temperature at the combustor exit, possibly requiring an enhanced structure as well as a fuel injection system, as stated in [4]. Results indicate that the performance of the topped engine is always higher than that of the corresponding baseline engine with the same compressor pressure ratio value. The benefit is greater for lower compressor pressure ratios, whereas for higher ones the benefit diminishes. In fact, for compressor pressure ratios greater than 20, almost no benefit can be obtained. This clearly favors the wave- rotor-topping for small gas turbines with low compressor pressure ratios, as was concluded in [5].

   A study aiming to investigate the performance trends of several four-port wave rotor topping configurations was done in [6]. Multi- parametric performance maps were generated by varying component efficiencies, pressure ratios, and temperatures. Even though wave rotor compression efficiency (WC) and expansion efficiency (WE) are functions of wave rotor temperature and pressure ratios, the above investigation assumed that WE =WC = 0.83. These values were also applied by other researchers in previous wave rotor studies, such as [7]. It was found that the engines efficiency is mainly affected by the polytropic turbine efficiency, PT, in comparison to the polytropic compressor efficiency, PC and combustion chamber pressure loss, Pcomb. Even though wave rotor topped gas turbine engines seems to be beneficial throughout a wide range of gas turbine sizes, efficiencies, and operating conditions, it was concluded in [6] that the most beneficiary ones are the small gas turbines with low component efficiencies and low temperatures.

   Wave rotor benefits were also assessed for the case of a two-spool short range turbofan engine at cruise conditions in [8]. For such engines, the wave rotor is suggested to be placed in the high pressure part of the engine, between the high pressure compressor (HPC) and the high pressure turbine (HPT), being in parallel to the combustion chamber.

   In [1], a recuperator was proposed to be placed downstream the wave rotor and before the combustion chamber. It was found that the baseline engine with recuperation has much higher efficiency – about twice – and lower specific fuel consumption than the ones of the unrecuperated baseline engine due to the succeeded reduction of heat input in the combustion process.

   The design of the transition ducts connecting the wave rotor to compressor, combustion chamber and turbine, seems to have a significant effect on the gas turbine performance. CFD studies were carried out in [9], considering the analysis of the subsonic, unsteady flow exiting the wave rotor with circumferential variation in total pressure and total temperature. Simulations showed that the flow at the exit of the transition ducting system has a total pressure loss of about 7% and a circumferential non-uniformity of total pressure. The latter is expected to influence the turbine performance connected to the wave rotor.

   From the literature survey presented above, it is concluded that many efforts have been made to exploit performance benefits when the baseline engine is topped with a four port wave rotor for aeronautical applications. Yet, ground gas turbines for power generation performance benefits have not been analyzed in detail. This study is an original contribution on gas turbine cycle calculations of one-shaft and two shaft regenerated gas turbines with heat exchanger, topped with a four port wave rotor. A performance map comparison between the regenerated topped engines at design point and the baseline regenerated engines provides the data to identify the types of engines and their operating characteristics that benefit mostly of possible integration of a four port wave rotor.

2. REGENERATIVE GAS TURBINE CYCLES CALCULATIONS

   1. Gas Turbine Regenerators

      The technique of adding a regenerator to the basic gas turbine cycle isbecoming prominent in our days of high fuel costs and tight reserves. The addition of heat exchanger is widely used in ground based power plants and its benefits in reducing fuel consumption and increasing power output are presented in [10]. Using regeneration, part of the heat available in the exhaust gas stream, otherwise lost, is recovered to preheat air from the compressor outlet before it enters into the combustion chamber. Considering using a heat exchanger with minimum pressure losses and achieving a negligible reduction of power output, the engine requires less fuel consumption thus increasing its thermal efficiency, as shown in [11]. The incorporation of a heat exchanger leads to higher efficiency at low pressure ratios. Additionally, the optimum pressure ratio for the recuperated cycle is much lower than that of the basic. The optimum pressure ratio for maximum specific work is sited between these two pressure ratios according to [12].

      However the presence of a heat exchanger in gas turbines for aeronautical applications is prohibited due to its weight, size and flow pattern complexity. In fact the only possibility to fit a heat exchanger in an aircraft engine is to use an air-to-gas compact type. Obviously this problem is obsolete in a ground based plant, where there is no constraint in size and weight and the coolant could be a liquid, [11].

      A lot of research efforts have been made on the design of a heat exchanger layout, including material selection, manufacturing technology, and optimization. Based on results reported in the past, a review paper in [13] provides potential heat exchanger designs for gas turbine engines.

   2. Input data

      Simple one-shaft gas turbine cycle calculations including a four port wave rotor were presented in detail by [14], so the calculation procedure will not be repeated here. It is important to mention that the calculations of the baseline as well as of the regenerated engine were performed assuming irreversible processes in all the gas turbine components. The thermodynamic model accounts also for cooling the turbine in case TIT>1300 K by subtracting air flow from the compressor. Typical values of input data are presented in Table I below. For all calculations it was assumed that the values of specific heat at constant pressure and ratios of specific heats are for the air Cpc 1005 J / KgK, c 1.4 and for exhaust gases Cph 1150 J / KgK, h 1.333

      TABLE I. BASELINE ENGINE TYPICAL INPUT DATA

      Quantity	Symbol, Unit	Value
      Ambient pressure	Pa , KPa	101.3
      Ambient temperature	Ta , K	288
      Intake pressure losses	Pin %	1
      Compressor pressure ratio	Rc	5 30
      Combustion chamber pressure losses	Pcc %	5
      Fuel Low Calorific Value	FCV, MJ/Kg	48.6
      Turbine Inlet Temperature	TIT, K	1000 1600
      Isentropic compressor efficiency	isc	0.85
      Combustion chamber efficiency	cc	0.99
      Isentropic turbine efficiency	ist	0.90
      Heat exchanger efficiency	nHE	0.88

   3. Wave Rotor thermodynamic calculations

      Typical input data for wave rotor thermodynamic calculations are summarized in Table II.

      TABLE II. WAVE ROTOR TYPICAL INPUT DATA

      Quantity	Symbol, Unit	Value
      Wave rotor pressure ratio	PR	1.4 2.2
      Ducting and leakage losses	Pduct %	8
      Efficiency of compression processes inside the wave rotor	nWC	0.75 0.92
      Efficiency of expansion processes inside the wave rotor	nWE	0.75 0.92

      Figure 2 illustrates the model used for the thermodynamic calculations of the four-port wave rotor. In this figure it is proposed to place the wave rotor in such a way that its cold inflow port (point 5.0) receives compressed air from the compressor exit and releases it after being further compressed to the heat exchanger cold inlet (point 5.1). The compressed air is preheated inside the regenerator and then directed to the combustion chamber (point 5.2). The hot gases from the combustion chamber enter the wave rotor hot inflow port (point 5.3), come in contact with the cold air being already inside the rotor channels and create compression waves to the air flow, get expanded

      Fig. 2. Symbols used for the four-port wave rotor thermodynamic calculations

      and finally directed to the outflow port towards the turbine (point 5.4). The heat exchangers hot inlet receives the hot gasses from the turbine outflow, enables heat exchange between the compressed air and the hot gases and releases them to the exhaust of the engine at a lower temperature.

      The calculation procedure is the following:

      Stagnation temperature at the cold air port of the wave rotor, T5.0

      Fig. 3. Four-port wave rotor integrated to a regenerated one-shaft industrial

      gas turbine

   4. One-shaft Gas Turbines with Heat Exchanger downstream the wave rotor

      One-shaft gas turbines are most suited for fixed speed operation such as base-load power generation. One-shaft gas turbines have the advantage of preventing over-speed conditions due to the high

      T5.0 T05

      (1)

      power required by the compressor and can act as an effective brake in case of electrical load loss, as stated in [10].

      It is assumed that the exit temperature from the outflow port of the wave rotor, T5.4 , equals the turbine inlet temperature, TIT

      T5.4 TIT (2)

      Stagnation pressure at the cold air port of the wave rotor P

      Figure 3 presents a schematic diagram of a one-shaft gas turbine with heat exchanger including a four-port wave rotor. In this configuration, the heat exchanger transfers heat from the hot gases at the exhaust of the turbine (point 8) to the compressed air exiting the wave rotor (point 5.1), prior to entering the combustion chamber (point 5.2).

      P P

      Pduct

      5.0

      (3)

      Figure 4 shows the corresponding Temperature – Entropy (T-S)

      5.0 05 1

      100

      thermodynamic diagram of the baseline engine as well as the wave rotor topped regenerated one. Points 2-3-4-7-8 correspond to the baseline engine. Points 2-3-4-5-5.1-5.3-5.4-8 correspond to the

      Stagnation pressure at the heat exchanger cold part

      P5.1 P5.0 PR

      P5.1

      (4)

      topped engine.

      Performance curves of various one-shaft regenerated engines at design point conditions without wave rotor (dotted lines) and topped with wave rotor (solid lines) are illustrated in Fig. 5, according to

      The stagnation pressure at the outlet port of the wave rotor towards the configuration shown in Fig. 3. Each symbol in Fig. 5

      the turbine,

      P5.4 , is obtained, assuming a linear variation of the

      corresponds to a different engine operating at design point

      P

      T

      conditions. The abscissa of

      pressure ratio

      5.4

      P

      versus the temperature ratio

      5.4

      T

      5.0 5.0

      across the wave rotor, based on experimental and numerical data presented in [15].

      Stagnation temperature at the wave rotor exit towards the heat exchanger, T5.1

      PR c 1. c 1

      T5.1 T5.0

      nWC

      1

      (5)

      Stgnation pressure at the combustion chamber outlet P5.3

      P5.3

      P5.2

      Pcc

      100

      100

      1

      1

      (6)

      Stagnation pressure at the hot inflow port of the wave rotor, T5.3

      T5.3

      P

      T5.4

      h 1/ h

      (7)

      Fig. 4. Thermodynamic diagram temperature-entropy for one-shaft baseline and topped gas turbine engine. Points 2-3-4-7-8 correspond to the

      1 1 5.4 nWE

      P5.3

      baseline engine. Points 2-3-4-5-5.1-5.3-5.4-8 correspond to the topped engine.

      the diagram is the specific work ( ws ) and the ordinate is the specific fuel consumption (sfc). These curves are obtained by varying the compressor pressure ratio (Rc) from 5 to 30 and the

      Turbine Inlet Temperature (TIT) from 1000 K to 1600 K.

      Figures 6 and 7 present a qualitative evaluation of the results obtained in Fig. 5. Figure 6 presents the percentage increase in specific work, ws , defined as

      ws,topped ws,baseline

      Comparing the performance curves of the baseline regenerated engines (without wave rotor) and the topped regenerated engines with wave rotor, for every single value of TIT, one can observe a

      % Increase in ws 100

      w

      w

      s,baseline

      (8)

      significant reduction in sfc and an increase in ws . The general trend of the engines topped with wave rotors is that they operate at lower

      values of sfc and at higher values of w . Thus, the performance of

      for different values of the compressor pressure ratio (being the abscissa of the diagram) and the turbine inlet temperature.

      From this figure it can be seen that for each value of Rc, the increase

      s

      the topped engines (continuous lines) is shifted to the right lower part of the diagram. It can be verified as observed in [14], that the shape of curves corresponding to values of TIT > 1200 K resembles to fish hook shape, except for TIT=1000 K and TIT=1100 K, where the curves are widened. Another observation is that for temperatures TIT=1500 K and TIT=1600 K and for compressor pressure ratios (Rc) greater than 15, sfc is remaining nearly constant.

      in ws is larger for engines operating at high values of TIT. The maximum gain in ws , as defined previously, for an engine with Rc=5 and TIT=1000 K is 12.6%, while for an engine operating at

      the same Rc and TIT=1600 K it is becoming 19.4%. For low values

      of Rc, there is a greater range of engines with gain in ws , operating with TIT from 1000 K to 1600 K. The gain decreases as the Rc increases and disappears at low TIT values.

      Fig.5. Performance of base line and topped with wave rotor one-shaft gas turbines at design point with heat exchanger, for wave rotor pressure ratio PR=2.2

      Fig. 6. Gain in specific work between base line and topped with wave rotor one-shaft gas turbines at design point with heat exchanger, for wave

      rotor pressure ratio equal to 2.2

      The percentage reduction in sfc, is defined as

      sfcbaseline sfctopped

      Fig. 8. Four-port wave rotor integrated to a regenerated single shaft industrial

      gas turbine

      As a result there is an air enthalpy increase providing an extra compression before being directed towards the combustion chamber of the gas turbine (point 5.1). Simultaneously, the hot gases inside the rotor are expanded (due to their contact with the air), which results to their enthalpy decrease as they are directed towards the turbine for a further expansion (at point 5.3). At the turbine exit, the

      % Reduction in sfc

      sfcbaseline

      100

      (9)

      pressure of the exhaust gases drops, but their temperature is high enough to exchange heat with the compressed air coming out from

      Figure 7 shows an important reduction in sfc, which is higher at

      lower values of TIT. For Rc=5 and TIT=1000 K the reduction in sfc is 54% compared to the untopped baseline engine with heat exchanger. At high values of Rc, engines with reduction in sfc, can operate only at high values of TIT. For example, when Rc=15, engines can operate only when TIT is greater than 1300 K. For this case, the reduction in sfc is 34.8%. Generally, for a given value of Rc, the largest reduction in sfc corresponds to the smallest possible value of TIT that the engine can operate with a benefit. This is why reduction in sfc for engines with Rc > 15 corresponds only to high values of TIT.

   5. One-shaft Gas Turbines with Heat Exchanger upstream the wave rotor

      The configuration shown in Fig. 8 is an alternative layout with respect to the positions of the wave rotor and the regenerator in the topped engine assembly. The air at the compressor outlet is preheated at the cold part of the regenerator (points 4 and 5) before entering the wave rotor (point 5.0). There, it comes in contact with the hot gases exiting the combustion chamber (point 5.2) and exchanges energy with them.

      Fig. 7. Reduction in specific fuel consumption between base line and topped with wave rotor one-shaft gas turbines at design point with heat

      exchanger, for wave rotor pressure ratio equal to 2.2

      the compressor exit when they enter the hot part of the heat exchanger (point 8).

      The thermodynamic T-S diagram of the baseline engine and the topped one is illustrated in Fig. 9. In this figure the main assumption is that both baseline and topped engines operate under the same compressor pressure ratio and the same turbine inlet temperature (TIT). The temperature rise from point 4 to point 5 is due to the heat exchanger. The numbers shown in Fig. 9 correspond to the locations shown in Fig. 8.

      Figure 10 illustrates performance curves of various one-shaft engines with heat exchanger according to the configuration shown in Fig. 8, at design point operation topped with wave rotor (solid lines) and without wave rotor (dotted lines). The abscissa of the diagram is the specific work ( ws ) and the ordinate is the specific fuel consumption (sfc). These curves are obtained by varying the compressor pressure ratio from 5 to 30 and the TIT from 1000 K to 1600 K.

      Fig. 9. Thermodynamic diagram temperature-entropy for one-shaft baseline and topped gas turbine engine. Points 2-3-4-7-8 correspond to the baseline engine. Points 2-3-4-5-5.1-5.2-5.3-8 correspond to the topped engine.

      Fig. 10. Performance of base line and topped with wave rotor one-shaft gas turbines at design point with heat exchanger, for wave rotor pressure ratio PR=2.2

      Comparing the performance curves of the baseline engines (without wave rotor) and the topped engines with wave rotor, for every single value of TIT, one can observe a reduction in sfc and an increase in

      ws for the latter ones. The observations from this figure are similar

      to the observations of Fig. 5, but this configuration shows less significant benefits in terms of sfc reduction.

      Figures 11 and 12 present a qualitative evaluation of the results obtained in Fig. 10. Figure 11 presents the percentage increase in ws for different values of the compressor pressure ratio (being the abscissa of the diagram) and the turbine inlet temperature. For reasons of comparison the same scale was used in both Figs. 6 and

      11. It can be observed that for each value of Rc, the increase in ws

      is larger at low values of TIT. The maximum gain in

      ws , as

      defined previously is 11.1% and it corresponds to Rc=10 and TIT=1000 K. At higher compressor pressure ratios than 10, the gain is decreasing, yet, the highest gain corresponds to the smallest possible operating TIT value.

      From Fig. 12, it can be seen a reduction in sfc, which is bigger at lower values of TIT. For reasons of comparison the same scale was used both in Figs. 7 and 12.

      Fig. 11. Gain in specific work between base line and topped with wave rotor one-shaft gas turbines at design point with heat exchanger, for wave rotor pressure ratio equal to 2.2

      Fig. 12. Reduction in specific fuel consumption between baseline and topped with wave rotor one-shaft gas turbines at design point with

      heat exchanger, for wave rotor pressure ratio equal to 2.2

      These figures show the same trend, but in Fig. 12, the reduction in sfc is less than half with respect to the previous configuration. For Rc=5, Fig. 12 shows reduction of 21.3% in sfc with respect to the baseline engine with heat exchanger at TIT=1000 K, whereas in Fig. 7 the reduction in sfc was 54%. From a given value of Rc, the largest reduction corresponds to the smallest possible operating value of TIT. As the compressor pressure ratio is increasing, the reduction in sfc is even lower with respect to the previous configuration.

   6. Two-shaft Gas Turbines with Heat Exchanger downstream the wave rotor

   Two-shaft gas turbines are also employed in power generation with the power turbine designed to operate at a fixed speed determined by the generator. Unlike one-shaft engines, the gas generator speed varies with electrical load.

   Fig. 13. Four-port wave rotor integrated to a regenerated two-shaft industrial

   gas turbine

   The main advantages with respect to one-shaft engines are the need for lower starting power requirements and better off-design performance. Yet, the disadvantage is that the shedding of the electrical load can result in over-speeding of the power turbine. A layout of a two-shaft regenerated engine, topped with a four port wave rotor is illustrated in Fig. 13.

   Figure 14 shows the performance map of regenerated two-shaft gas turbines at design point. Continuous lines depict the results obtained for the topped engines, while dashed lines depict the results of the baseline engines. The input data used – Tables 1 and 2 – is the same as the one for the case of one-shaft regenerated engines (Fig. 5). By comparing Figs. 5 and 14, it can be observed that both types of engines have almost the same performance. However, it must be

   noted that the performance is slightly better both in terms of ws

   and in sfc in the case of the two-shaft engines.

   Fig.14. Performance of base line and topped with wave rotor two-shaft gas turbines at design point with heat exchanger, for wave rotor pressure ratio PR=2.2

3. CONCLUSIONS

In this paper performance benefits for regenerated ground gas turbine engines topped with four-port wave rotors were investigated. Calculations were performed for two typical categories of engines. The first category is the one-shaft engines, where two possible positions of placing the wave rotor were examined. According to the first one, the wave rotor is placed between the regenerator and the combustion chamber. Thermodynamic analysis of the regenerated and topped cycle showed benefits both in terms of gain in specific

work,

ws , and in reduction of specific fuel consumption, sfc. The

Fig. 15. Gain in specific work between base line and topped with wave rotor two-shaft gas turbines at design point with heat exchanger, for wave rotor pressure ratio equal to 2.2

Figure 15 shows the percentage increase in ws with respect to the baseline regenerated engines. As for the case of one-shaft engines it can be observed that for each value of Rc, the increase in ws is larger at high values of TIT. The maximum gain in ws , as defined by Eq. (8), is 14.7% when Rc=5 and TIT=1000 K (instead of

12.6% for the case of one-shaft topped regenerated engines) , while

at Rc=5 and TIT=1600 K it is becoming 20.5% (instead of 19.4% for the case of one-shaft topped regenerated engines). For low values of Rc, there is a greater range of engines with gain in ws , operating with TIT from 1000 K to 1600 K. The gain decreases as the Rc increases and disappears at low TIT values.

From Fig. 16, it can be seen an important reduction in sfc, which for

a given compressor pressure ratio, Rc, is bigger at lower values of TIT. For Rc=5 and TIT=1000 K the reduction in sfc, as defined by Eq. (9) is 56.4% (instead of 54% in the case of one-shaft topped regenerated engines). At high values of Rc, engines can operate at design point only at high values of TIT. For example, when Rc=15, solution can be acceptable only when TIT is greater than 1300 K. For this case, the reduction in sfc is 38.5% (instead of 34.8% in the case of one-shaft topped regenerated engines).

Generally, for a given value of Rc, the largest reduction in sfc corresponds to the smallest possible value of TIT that the engine can operate with a benefit. This is why results for cases that Rc > 15 correspond only to high values of TIT.

Fig. 16. Reduction in specific fuel consumption between base line and topped with wave rotor two-shaft gas turbines at design point with heat exchanger, for wave rotor pressure ratio equal to 2.2

second possible position is to place the wave rotor at the exit of the compressor followed by the regenerator. This layout showed significant improvement in the performance of the topped engine and is the one proposed for implementation.

Having finalized the most advantageous position of the wave rotor in the regenerated one-shaft engine, calculations were also performed for two-shaft engines at design point. Once more an enhanced performance compared to the baseline engines was verified by the calculations.

The gain in specific work is maximized for engines operating at low values of compressor pressure ratios and high values of turbine inlet temperatures. The reduction in specific fuel consumption is maximized for engines operating at low values of both compressor pressure ratios and turbine inlet temperatures.

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