Effect of combustion chamber insulation on the performance of a low heat rejection diesel engine with exhaust heat recovery

Heat Recover.)"Systems & CHP Vol. 9, No. 5, pp. 475-484, 1989

0890-4332/89 $3.00 + .00 Pergamon Press plc

Printed in Great Britain

EFFECT OF COMBUSTION C H A M B E R I N S U L A T I O N ON THE P E R F O R M A N C E OF A LOW HEAT REJECTION DIESEL E N G I N E WITH EXHAUST HEAT RECOVERY D. N. ASSANIS Department of Mechanical and Industrial Engineering, 1206 W. Green Street, University of Illinois at Urbana-Champaign, Urbana, IL 61801, U.S.A. (Received 6 March 1989)

Almtraet--A computer simulation of the turbocharged turbocompound diesel engine system is used to study the effect of combustion chamber insulation on the performance of low heat rejection system configurations with exhaust heat recovery. The analysis is carried out for zirconia coatings of various thicknesses applied on the cylinder head and piston. It is found that an intercooled turbocompound engine derives a modest thermal efficiency benefit from insulation, e.g. 4.3% improvement at a 60% reduction in heat loss. The addition of Rankine compounding can improve the thermal efficiency of the turbocompound engine by 10--14%, depending on the level of insulation and the system configuration. Furthermore, Rankine compounding can make the otherwise inferior performance of a non-intercooled engine match the performance of an intercooled engine. Fina!l,v, use of an insulating material of low conductivity and low heat capacity can increase the thermal efficiency benefits, but at the expense of increased component thermal loading.

INTRODUCTION The use of ceramic materials in diesel engine component design is especially promising. It has been shown that significant reductions in cooling and exhaust system losses can be obtained by lining critical components of the engine with ceramics [1]. One of the most practical concepts for taking advantage of this extra available energy is the turbocharged turbocompound diesel engine system, shown in Fig. 1. This engine concept consists of several subsystems: compressor, intercooler, intake manifold, reciprocator, exhaust manifold, turbocharger turbine, compound turbine and ducting. A computer simulation of the complete system has been developed to enable engineers to define the thermal efficiency benefits and the trade-offs associated with the use of ceramics in different system components. Details of the model development and validation can be found in [2]. The present work focuses on the effects of combustion chamber insulation on the performance of alternative low heat rejection system configurations. One objective is to explore the relationship between the thickness of insulation, degree of reduction in heat loss and resulting system performance improvement. This is important in determining which surfaces to insulate and to what extent, and also, in defining the characteristics of ideal insulation materials and the thermal environment that they must withstand. A second objective is to examine which low heat rejection system configurations are more desirable for different levels of insulation. Examples of such configurations include turbocharged turbocompound systems with or without intercooling, and with or without additional exhaust heat recovery devices such as Rankine bottoming systems. These issues are explored through the systematic use of the computer simulation which is briefly summarized below. SUMMARY OF SYSTEM SIMULATION In the computer simulation, the reciprocator cylinders, the intake manifold and the various sections of the exhaust manifold are treated as a series of connected open systems. The contents of each of these systems, i.e. air and combustion products, are represented as one continuous medium by defining an average, equivalence ratio and a uniform temperature and pressure at all times. Gas properties arc calculated assuming ideal gas behavior, with allowance for chemical dissociation at high temperatures. 475

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The diesel four-stroke cycle is treated as a sequence of continuous processes: intake, compression, combustion (including expansion) and exhaust. Quasi-steady, adiabatic, one-dimensional flow equations are used to predict mass flows past the intake and exhaust valves. Combustion is modeled as a uniformly distributed heat release process. The sum of two algebraic functions, one for the pre-mixed and the other for the diffusion-controlled combustion phase, is used to describe the rate of fuel burning [3]. The relative weight of each combustion phase depends on the length of the ignition delay period and the engine load and speed. Heat transfer is included in all the engine processes. Convective heat transfer is modeled using available engine correlations based on turbulent flow in pipes. The characteristic velocity and length scales required to evaluate-these correlations are obtained from a mean and turbulent kinetic energy model. Radiative heat transfer, based on the predicted flame temperature is added during combustion. The time-dependent temperature distribution in the piston, cylinder head, liner and manifold walls can be computed using the transient conduction models which are described in the next section. An empirical friction model is used to convert indicated to brake engine performance quantities [4]. The turbomachinery performance is defined by maps that interrelate elticiency, pressure ratio, mass flow rate and shaft speed for each component. The compressor, turbine and compound turbine are assumed to be adiabatic. An open-system duct model connects the turbocharger turbine with the compound turbine. The turbocharger dynamics are controlled by the rotor inertia and damping. Finally, the compound turbine shaft is connected to the engine crankshaft via a specified gear ratio transmission. WALL HEAT C O N D U C T I O N MODELS The heat transfer rates from the gas to the walls of the various system components depend on the instantaneous difference between the gas and the wall temperature. For most components, temperature varies very rapidly in directions perpendicular to the surface, so that heat transfer by conduction through the walls can be modeled as one-dimensional. Furthermore, for engines with high conductivity metal walls and forced convection jacket cooling, measurements by LeFeuvre [5] suggest that cyclic surface temperature variations are small, ranging from 5 to 15 K. However, for engine surfaces insulated with low conductivity and low diffusivity materials such as ceramics, these

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surface temperature variations are most critical [2]. In order to predict the time-dependent temperature distribution in each component, we then proceed as follows. First, the steady-state temperature distribution is obtained through an iterative procedure. At the start of the cycle simulation, estimates of the steady-state inside wall surface temperatures are assumed. Based on these estimates, the instantaneous heat transfer rates, convective and radiative, to the combustion chamber walls are calculated throughout the engine operating cycle. At the end of the cycle, a heat balance is performed between the cycle-averaged gas to wall heat transfer rate and the heat conducted through the walls of each component to compute new surface temperatures. These new temperatures are used in the next engine cycle iteration until the calculation converges. Once the steady-state temperature distribution within each component wall is obtained, a suitable numerical technique is used to solve the transient conduction problem. Each component is modeled as a number of discrete nodal points. At each node, the instantaneous temperature is related to those of the surrounding nodes through a finite difference approximation of the governing conduction equation. Special effort was made to develop a numerical scheme that would be: (i) least demanding in computer time; (ii) able to handle arbitrarily-spaced discrete nodes so as to maximize the accuracy of the solution within the relatively thin penetration depth of the cyclic engine transients, and (iii) stable. Details of the scheme can be found in [2]. M E T H O D OF SOLUTION The conservation equations of mass and energy for the contents of an open thermodynamic system are applied in turn to the master reciprocator cylinder, the intake manifold and the series of exhaust manifold sections. Further, the individual submodels of the various system components and their thermodynamic and heat transfer processes are brought together to form a complete system model. The result is a set of simultaneous first-order ordinary differential equations which are integrated numerically using a predictor-corrector technique. The input parameters which must be specified for each cycle simulation calculation include system operating conditions, system dimensions and design parameters, wall structure specifications, initial conditions of the system and certain computational parameters. The output includes mean engine performance parameters, such as power, specific fuel consumption, mean effective pressure and thermal efficiency, as well as detailed information about the state of the total system as a function of crank-angle throughout each engine cycle. To validate the diesel system model, simulation predictions have been compared against data obtained from Cummins for a six-cylinder, 14.0 liter, cooled, turbocompound diesel engine. Figure 2 shows predicted and measured reciprocator brake power, boost pressure ratio and overall brake specific fuel consumption as a function of reciprocator speed for operation at constant load. The variation of all three predicted performance quantities closely follows the data in both magnitude (within 2.5%) and trend. The small discrepancies (primarily at low speeds) can be attributed to inadequacies in the heat transfer models, possible errors in the assumed variation of friction with speed and limitations in the range of the turbomachinery maps. Good agreement between measured and predicted performance of the baseline turbocompound engine was also obtained over a range of loads for a given reciprocator speed [2]. Hence, with the above qualifications, the model gives sufficiently accurate predictions of trends and magnitudes to be useful for assessing the effect of changes in system design and operating variables on system performance. E F F E C T OF C O M B U S T I O N C H A M B E R I N S U L A T I O N ON SYSTEM P E R F O R M A N C E Having summarized the model assumptions, we can now use the simulation to examine the effect of various degrees of combustion chamber insulation on system performance. The relationship between thickness of insulating material and resulting degree of reduction in heat loss is studied for a series of coatings of sprayed zirconia (conductivity k -- 1.2 W m-~ K-~ ) applied on a l0 mm thick cast iron cylinder head and piston. For an optimum insulation strategy [6], the 7 mm cast iron liner is maintained cooled. For all the combustion chamber components, the wall surface HR.S. 9.~--F

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temperatures on the cold side are set at 380 K. All performance predictions are carried out at the rcciprocator's rated speed (1900 rpm) and a fixed fueling rate (17.9 g s-'). The compression ratio is held constant at 14.5. The compound gear ratio is fixed at 18: l, while gearbox transmission losses are assumed to be 5% of compound turbine power. Figure 3 (top) shows the surface temperature profiles for different material thicknesses of sprayed zirconia, ranging from l to 8 mm. As thickness increases, the mean temperature at the cylinder head and piston surfaces increases (from 650 to 920 K), while the gas to wall heat transfer is reduced. The other characteristics of these profiles (shape, peak-to-peak amplitude) are similar since the Y

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swings around the mean surface temperatures are essentially constant with insulating thickness (Fig. 3, bottom). It is interesting to note that these swings are substantially greater than thbse obtained for a baseline, cooled, cast-iron piston. Figure 4 shows graphically the relationship between the insulating thickness of zirconia and the resulting reduction in heat loss to the coolant. The rapid reduction in heat loss with an initial increase of thickness must be noted; a 3 mm insulating layer is sufficient to cut down heat losses by 43%. As more insulation is added, the incremental benefit progressively diminishes; for instance, an additional 7 mm of insulation reduces losses by an extra 15% only. The reason is that the fully insulated condition for the piston and the cylinder head is approached. The rest of the heat losses (i.e. approximately 40%) are associated with the non-insulated liner. The above results are in good agreement with previously published studies [6, 7]. Figure 5 summarizes the effect of combustion chamber insulation on the performance of an intercooled, turbocompound engine (solid lines). Also shown in Fig. 5 are corresponding results for a non-intercooled version of this engine (dashed lines); the latter are discussed in a subsequent section. One of the important conclusions is that reductions in heat loss due to insulation are primarily traded for increased exhaust enthalpy. This is reflected in the substantial rise of the mass-averaged temperature of the exhaust gas shown in Figs 5a, b. In contrast, there is little change in diesel engine power (see Fig. 6). This result can be explained in terms of the degradation in volumetric efficiency (based on intake manifold conditions) that occurs with an increasing degree of insulation (Fig. 5c). This drop results from the decrease in charge density due to the heat transferred to the intake charge from the high temperature insulated walls. Since the engine is turbocharged, part of the extra available exhaust energy is used by the turbocharger turbine which supplies more power. The resulting higher boost presure ratios (Fig. 5d) compensate for the reduction in charge density so that the intake air flow remains constant with insulation (Fig. 5e). Nevertheless, the intake charge starts from a higher temperature and higher pressure (due to the higher boost pressure ratios). As a result, more compression work is required in an insulated cylinder compared to the baseline engine. Since the injection timing and the compression ratio are kept constant, the increased boost pressure and the reduced ignition delay result in increased peak cylinder pressures (Fig. 50. However, the substantial gas to wall heat transfer during combustion (despite the insulation) limits the potential improvement in expansion work. Assuming that engine friction does not vary with insulation, the little change in compression/expansion work results in a relatively small improvement (3 to 4%) in cylinder brake thermal efficiency (Fig. 5g). The loss in volumetric efficiency affects the performance of the compound turbine, too, since most of the available high pressure and high temperature energy at the engine exhaust is used-up by the turbocharger turbine to provide the higher boost levels. The remaining high temperature, but low pressure energy at the inlet of the compound turbine is only partially converted into useful work as shown by the steadily increasing temperatures at the exhaust of the compound turbine (Fig. 5b). Since both the reciprocator and the compound turbine performance gains are modest, the overall brake thermal efficiency improvement over a baseline turbocompound engine is 4.3% at a 60% reduction in heat loss (Fig. 5g).

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OPTIMIZATION

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The above results suggest that, at rated conditions, the performance gains associated with practical levels of insulation are fairly modest. However, a comprehensive optimization process needs to be carried out before drawing any final conclusions. This paper concentrates on the optimization of system configuration and selection of insulating material.

Optimization of system configuration lntercooling. The analysis of the previous section has been repeated for a turbocompound diesel engine without intercooling. Although the results are similar in trend to those obtained for the intercooled engine (see Fig. 5), there are some important differences, too. First, the volumetric efficiency of the non-intercooled engine is consistently higher (about 5%) than that of the intercooled engine for all levels of insulation. This is attributed to the reduced heat transfer during

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the intake process which is caused by the higher (by about 90 K) temperatures in the intake manifold of the non-intercooled engine. As a result, gas temperatures throughout the engine cycle increase and this leads to increased available energy at the engine exhaust (Fig. 5a). However, despite the resulting higher boost pressure ratios, the turbocharger can only partially compensate for the dramatic reduction in charge density due to the removal of intercooling. Consequently, the intake mass flow is reduced substantially compared to the baseline intercooled engine (Fig. 5e). Furthermore, the higher gas temperatures during the cycle result in higher levels of in-cylinder heat transfer and reduced peak cylinder pressures; this leads to about 2% lower power levels (Fig. 6) and 1% lower thermal efficiency levels (Fig. 5g) compared to the corresponding intercooled engine configurations. It can be concluded that intercooling the turbocompound diesel is beneficial to system performance for all levels of insulation. Rankine Compounding. Insulation and/or removal of intercooling results in increased available energy in the form of high temperatures at the exhaust of the compound turbine. Other exhaust heat recovery systems such as Rankine compounding have been considered in addition to turbocompounding [8] to utilize more of this energy. A schematic of a diesel engine compound with a Rankine bottoming cycle system is shown in Fig. 7. Simulation and testing of such Rankine bottoming cycle systems operating with steam [9] and organic working fluids [10] have shown that they can convert 17-24% of the heat extracted from the exhaust gas into useful power, depending on the engine and the Rankine cycle operating conditions. These calculations are based on a minimum exhaust stack temperature of 422 K (sulfuric acid condensation limit), a maximum turbine pressure of 7 MPa and a negligible exhaust pressure drop across the boiler. Figure 6 (bottom) shows the additional power than can be extracted from a Rankine bottoming system for both an intercooled and a non-intercooled version of the turbocompound engine. These calculations are based on a Rankine cycle efficiency of 21%. Two important conclusions can be drawn. First the power produced by the Rankine compounding is substantial and increases with insulation at a faster rate than the compound turbine power; thus Rankine compounding can improve the performance of the turbocompound engine by 10-14% depending on the level of insulation. Second, the non-intercooled engine benefits more from the addition of Rankine compounding than the intercooled engine, due to its higher temperatures at the compound turbine exhaust (Fig. 5b). As a result, the non-intercooled engine with turbine plus Rankine compounding can match the performance of the corresponding intercooled system configuration. Of course, these benefits of Rankine compounding must be evaluated carefully against the added complexity, cost, maintenance, reliability and packaging difficulties of the Rankine system.

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Selection of insulating material A wide range of ceramic materials and methods of application have been proposed for insulated engine components. In order to illustrate the engineering trade-offs associated with the use of alternative ceramic materials to achieve a desired degree of insulation, the simulation was used to study: (a) A 1.5 mm plasma-sprayed zirconia coating of conductivity k = 0.6 W m- ' K- ' and heat capacity pc = 1.1 x 10 6 J m -3 K - I [11] applied to a 10 mm cast iron cylinder head and piston; (b) A 5 mm Partially-Stabilized Zirconia (k = 2 W m - ' K - ' and pc = 3.14 x 106 J m-3 K - ' insert in a 10mm cast iron cylinder head and piston; (c) A 13.4mm monolithic Reaction Bonded Silicon Nitride (k = 5 W m-' K - ' and pc = 1.78 x l & J m -3 K - ' cylinder head and piston without a cast iron substructure; Obviously, the thickness of insulation was adjusted in each case to compensate for differences in conductivity so as to achieve a 43% reduction in heat transfer over the baseline cast iron engine. Also, the liner was non-insulated, 7 mm thick case iron, and the wall surface temperatures on the cold side were kept at 380 K for all components. Figure 8 shows the cyclic variations from the mean temperature (800 K) at the surface of the piston for the three candidate materials. Clearly, the plasma-sprayed zirconia, which has the lowest conductivity and lowest thermal capacity of the three materials, experiences the largest surface temperature swings (235 K peak-to-peak) around the mean temperature. In contrast, the swings predicted for the PSZ (95 K) and the RBSN (70 K) are substantially smaller. Since the larger the surface temperature swings, the greater the reduction in instantaneous gas-to-wall heat transfer rates, the zireonia-coated engine achieves a higher overall brake thermal efficiency (46%) compared 20C ,R 03

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cast iron piston Fig. 9. Transient temperature profiles for a zirconia-coated piston (top) and a baseline ba~ (bottom). to the PSZ (45.5%) and RBSN (45.4%) insulated engines. However, before deciding on an optimum material for a given application, it is important to consider the associated component thermal loading. Figure 9 compares the transient temperature profiles (i.e. perturbations from the steady-state temperature distributions) for a zirconia-insulated and a conventional cast iron piston at three instants during the engine cycle. Although the profiles are qualitatively similar, temperature and length scales are substantially different in these two cases. Due to the low thermal diffusivity of sprayed zirconia, the cyclic transients penetrate into the wall structure for only a limited distance (less than 0.5 mm). For comparison, the skin depth of the cast iron piston is 2.8 mm. The combination of large tcrnperature swings and short penetration depths in the zirconia-coated piston results in high thermal gradients, and thus high stress levels. We may conclude that the selection of an insulating material involves a strong trade-off between system performance improvement and component thermal loading. CONCLUSIONS This paper has described the application of a computer simulation of the turbocharged turboeompound diesel engine system to study the effect of combustion chamber insulation on the performance of various low heat rejection system configurations. The major conclusions are the following: (1) A relatively thin (3 mm) initial coating of sprayed zirconia can result in a substantial (43%) reduction in heat loss. However, as more insulation is added, the incremental benefits progressively decrease. (2) High combustion chamber wall surface temperatures due to insulation cause a significant drop in volumetric efficiency. Although this reduction in volumetric efficiency is offset by increased boost pressure from the turbocharger, an intercooled turbocompound engine can derive a modest thermal efficiency benefit from insulation (4.3% improvement at a 60% reduction in heat loss). (3) The intercooled, turbocompound engine maintains a steady performance advantage over the non-intercooled version of the engine for all levels of insulation examined. Apparently, turbocompounding cannot take advantage of the extra available energy duc to the removal of intercooling. (4) Addition of Rankine compounding improves the thermal efficiency of a turbocompound engine by I0-14%, depending on the level of insulation and the system configuration. This benefit and the possible elimination of intcrcooling must be traded against the added complexity of the overall system configuration.

484

D.N. ASSANIS

(5) F o r ceramic insulated c o m p o n e n t s , cyclic surface t e m p e r a t u r e variations are critical. The lower the thermal c o n d u c t i v i t y a n d the lower the thermal capacity of the material, the higher the wall surface t e m p e r a t u r e variations, the smaller the d e g r a d a t i o n in volumetric efficiency a n d thus the better the t h e r m a l efficiency of the overall system. However, the material must be able to w i t h s t a n d adverse t h e r m a l gradients. REFERENCES 1. R. Kamo and W. Bryzik, Adiabatic Turbocompound Engine Performance Predictions, SAE Paper 780068 (1978). 2. D. N. Assanis and J. B. Heywood, Development and use of a computer simulation of the turbocompound diesel system for engine performance and component heat transfer studies, Paper 860329, SAE Trans. 95, 451-476 (1986). 3. N. Watson, A. D. Pilley and M. Marzouk, A Combustion Correlation for Diesel Engine Simulation. SAE Paper 800029 (1980). 4. B. W. Millington and E. R. Hartles, Frictional losses in diesel engines, SAE Trans. 77, Paper 680590 (1968). 5. T. LeFeuvre, P. S. Myers and O. A. Uyehara, Experimental Instantaneous Heat Fluxes in a Diesel Engine and their Correlation, SAE Paper 690464 (1969). 6. W. Bryzik and R. Kamo, TACOM/Cummins Adiabatic Engine Program, SAE paper 830314 (1983). 7. V. Sudhakar, Performance Analysis of Adiabatic Engine, SAE Paper 840431 (1984). 8. N. Watson, N. P. Kyrtatos and K. Holmes, The performance potential of limited cooled diesel engines, 1. Mech. E. Proc. 197A, 197-207 (1983). 9. E. Poulin, R. Demler, L. Krepchin and D. Walker, Steam Bottoming Cycle for an Adiabatic Diesel Engine. DOE/NASA/0300-1, NASA CR-168255 (1984). 10. L. R. DiNanno, F. A. DiBella and M. D. Koplow, An RC-I Organic Rankine Bottoming Cycle for an Adiabatic Diesel Engine, NASA CR-168256 (1983). 11. J. D. Cawley, Overview of zirconia with respect to gas turbine~applications,NASA Technical Paper 2286. NASA Lewis Research Center (1984).