Applied Energy 152 (2015) 109–120
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Waste heat recovery of an ORC-based power unit in a turbocharged diesel engine propelling a light duty vehicle D. Di Battista ⇑, M. Mauriello, R. Cipollone Department of Industrial and Information Engineering and Economics, University of L’Aquila, Italy
h i g h l i g h t s Experimental campaign on a ORC-based power unit bottomed an IVECO F1C engine. Evaluation of the effect on the engine (backpressure increase). Evaluation of the effect on a light duty vehicle (weight increase). Investigation on the heat exchanger and expander off-design working conditions. Assessment of the net power improvement related to ORC unit on light duty vehicle.
a r t i c l e
i n f o
Article history: Received 21 February 2015 Received in revised form 20 April 2015 Accepted 21 April 2015 Available online 15 May 2015 Keywords: Waste heat recovery Turbocharged diesel engine Backpressure ORC Plate heat exchanger Specific fuel consumption
a b s t r a c t The present technologies in internal combustion engines for transportation purposes clearly demonstrate the room for improvement still achievable. In a recent past, harmful emission reduction was the main goal: wonderful technologies were developed which strongly reoriented the interest and the use of such engines. Actually, CO2 is the most important driver: it calls for fuel consumption reduction (energy saving) and energy recovery from that usually wasted. Considering that about one third of the fuel energy is in the flue gases, the possibility to recover this energy and re-use it for engine and vehicle needs is one of the smartest ways to participate to reduce fuel consumption and, therefore, CO2 emissions. In particular, Organic Rankine Cycle (ORC)-based power units fed by the exhaust gases are promising and technologically ready, but they have a significant impact on the exhaust line and engine behavior. A trade-off between energy recovered in mechanical form and energy lost due to the engine back pressure, vehicle weight increase, discharge energy at the condenser, and the management of the strong off-design operating conditions is a key point which could definitively open the way to this technology or limit it to particular applications. The paper discusses the effects of the pressure losses produced by an ORC-based power unit mounted on the exhaust line of a turbocharged IVECO F1C engine, operated on a test bench. The interactions produced on the turbocharged engine have been experimentally investigated: the presence of an Inlet Guide Vane (IGV) system to manage the turbocharger makes the effect of the back pressure not straightforward to be predicted. The IGV opening and closure degree, in fact, can compensate the effect of the back pressure which intrinsically tends to increase specific fuel consumption. A wide experimental testing of the turbocharging group in order to understand its reaction and the net effect in terms of engine specific fuel consumption is presented. Finally, once the engine performances were verified, the contribution due to the heat recovery inside the ORC-based power unit fed by the exhaust gases in terms of mechanical power was evaluated and experimentally verified in some points, considering the strong off design conditions produced by the engine operating point variations. In fact, exhaust gas flow rate and temperature variations lead the evaporator, even though properly designed, to severe off-design conditions which modify the inlet working fluid conditions till to make the mechanical recovery impossible. Under the hypothesis that the engine propels a light duty vehicle, the effect of the extra weight is discussed re-evaluating the propulsion power increase in terms of fuel consumption. Ó 2015 Elsevier Ltd. All rights reserved.
⇑ Corresponding author. E-mail addresses:
[email protected] (D. Di Battista),
[email protected] (M. Mauriello),
[email protected] (R. Cipollone). http://dx.doi.org/10.1016/j.apenergy.2015.04.088 0306-2619/Ó 2015 Elsevier Ltd. All rights reserved.
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Nomenclature A BSFC ECU EGR GWP H HRVG IGV N Nexp ODP OEM ORC P Qexch Qexhaust S VGT Vsuc h kc,1 kc,2 kc,3
cross section brake specific fuel consumption electronic control unit exhaust gas recirculating global warming potential enthalpy heat recovery vapor generator inlet guided vanes engine speed expander speed ozone depletion potential original equipment manufacturer organic rankine cycle power heat exchanged exhaust thermal power heat exchange surface variable geometry turbine expander suction volume heat transfer coefficient first BSFC calibration coefficient second BSFC calibration coefficient third BSFC calibration coefficient
1. Introduction Currently, the major global concern of the transportation sector on the road is the environmental aspect related to propulsion systems. In fact, fuel economy and emissions reduction are the ultimate reasons behind the actual technological research in this field, and global vehicle efficiency is one of the best ways to reach these goals. In particular, Euro regulation and carbon dioxide targets are constraining the recent and future developments in vehicle technologies. The first one is aimed to reduce pollutants (in particular particulate matter and nitrogen oxides), while the second one sets at 95 g/km by 2021, the CO2 emission level for passenger cars and light-duty vehicles (with a reference mass of 1392 kg); these limits reach 147 gCO2/km for vans (with reference mass of 1706 kg) by 2020. More recently, the European Parliament proposed a target of 68 gCO2/km by 2025. This is very challenging, considering that they have to be reached without modifying the traditional driving expectations (torque, power, acceleration, fun-to-drive, etc. . .) breaking the relationships between engine and environmental performances. Fig. 1 shows how the increase in engine power (also per displacement unit) has been accompanied by a fuel consumption improvement, leading to more efficient vehicles also keeping a good fun-to-drive feature (0-to-100 km/h time). The constraints on the CO2 emissions (specific fuel consumption) which are rapidly approaching, will push the technological advancement even further: they involve several engine and vehicle aspects. Each of them is characterized by a different carbon intensity benefit per unit of distance traveled (gCO2 saved/km). Among all the available technologies, the first ones which will appear on the market will have a lower specific cost increase per unit of CO2 saved (Euro/gCO2 saved) and minor on-board modifications on the engine and vehicle. Downsizing, transmission optimization, start and stop [2] are yet settled in vehicles; engine efficiency improvement [3,4], thermal management [5,6] and exhaust heat recovery [7,8] have reached a level of development to be almost market ready. Knowing that the penalties imposed
km kp kwall m mgas nvanes p t v Dp DT
g n
q
exhaust gas mass flow rate calibration coefficient pressure drop calibration coefficient thermal conductivity of the heat exchanger walls generic mass flow rate exhaust gas mass flow rate expander vanes number pressure thickness fluid velocity HRVG pressure drop temperature difference efficiency permeability scale factor density
Subscripts in inlet out outlet sl saturated liquid sv saturated vapor ORC organic rankine cycle WF working fluid (R236fa)
by the EU regulation is of 95 €/gCO2 exceeding the limits, a technology extra cost of about 60–70 €/gCO2 would be acceptable. This is the main reason why important technologies, like hybridizations, still have difficulties to be extensively introduced on the market. On the other hand, technological options like engine cooling optimization (30–50 €/gCO2), tires and transmissions optimization (30–40 €/gCO2) and waste heat recovery (60–80 €/gCO2) are considered acceptable and a stronger interest is reversed to them in order to achieve the international targets [9]. With reference to the heat recovery, it is known that the energy in the gases has a magnitude almost equal to the engine mechanical energy, representing a big challenge for recovery. Energy recovery performed with turbo-compounding [10–12] and with thermoelectric devices [13,14] are significant examples but they
Fig. 1. Engine performances trends in transportation sector [1].
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require important engine modifications or cost increase, today too high and the reference to technologies not completely proven. Turbo-compounding has been the subject of a strong interest because it is able to recover the enthalpy of gases due to a higher pressure as it is actually done to drive the compressor, [15,16]. For the intrinsic type of recovery, this technology can be addressed as ‘‘direct’’ recovery [17]. Main concern of this direct recovery is represented by very high speed of revolution of the turbine (due to the high flow rates) which partially prevents the mechanical transformation into electric energy or a direct use in mechanical form. The suitability is, therefore, limited to the mild or fully hybrid propulsion systems [18] and to new electric generator technologies. For these reasons, scientific and technical literature has concentrated a greater attention on the indirect recovery systems based on Organic Rankine Cycles (ORC) [19,20] power units. The high enthalpy of the exhaust gases (in terms of high temperature) is recovered inside an evaporator which produces a medium–high pressure vapor of an organic fluid. It is expanded converting the enthalpy in useful work; if expander is of a rotary volumetric type, revolution speeds are low and fully compatible with conventional electric generators or with a mechanical use. After the condensation of the exhausted vapor, it is re-pressurized by a pump. The advantages of an ORC-based unit for vehicle applications are outstanding. In fact, for many aspects the power unit could be downsized thanks to suitable technologies for the pump and the expander [21,22], even though the heat exchangers size still represent a crucial aspect and deserve specific developments. The cost increase can also be reduced thanks to component integration, leading to a value cheaper than other technologies. In literature, several studies have been concentrated on fluid choice [23,24] for which low ODP and GWP seem to be the most important constraints; suitable operating temperature range have been treated considering fluid degradation and component simplification. A great attention has been also focused on thermodynamic analysis of the cycle [25,26]: high efficiency is searched, including a regeneration phase and different thermodynamic options, like dual-loop or transcritical cycles [27,28]. Components choice and sizing are, also, investigated in literature: in particular, heat exchangers [29] and expanders [30,31] are the most worthy of care: their sizing is really important but, in particular in transportation sector, their resilience to off-design behavior appears as one of the most important aspects. Engine operations, in fact, are characterized by high flow rates and temperature variations which strongly influence the heat really recovered in the fluid or the mechanical conversion inside the expander. Dynamic modeling and optimal control strategy are usually used to predict waste heat recovery performances [32,33], but they still need to be further investigated. More recently, the integration of an ORC-based power unit with an internal combustion engine in the transportation sector has been presented [34]. Main drawbacks are represented by: (a) weight increase, which produces a propulsion energy increase which can make nil the energy recovered; (b) back pressure produced by the heat exchanger on the exhaust gases [35]; (c) the frequent off design conditions of all the components (mainly, expander, condenser and evaporator) which can sensibly reduce the energy recovered; (d) the heat discharge devices at the condenser. As per the authors, all these aspects have never been considered in a whole system analysis: it is certain that a trade-off between mechanical energy recovered and different drawbacks calls for a suitable choice of the design point and for an intelligent (model based) control. Single aspects, in reality, have been extensively treated. The effects of weight increase on propulsion power and heat exchanger and expanders off design conditions are matters of basic mechanical engineering. On the other hand, back pressure effects strongly influence engine performances [36,37]. Excessive
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back pressure in the exhaust system create excessive heat, turbocharger problems, increased pumping work, lower engine power and fuel penalty, that may cause damage of the engine parts and poor performance [38]. The amount of power lost due to back pressure depends on many factors, but a good rule-of-thumb is that 30 mbar back pressure causes about 1.0% loss of maximum engine power [39]. Recent studies on turbocharged engines assumed a fuel penalty of 2% having 100 mbar back pressure [40]. In a turbocharged diesel engine, as back pressure increases, the pressure ratio across the turbine decreases and the engine has to pump the gases out of the cylinder against a higher pressure, increasing the pumping. This would immediately produce a fuel consumption increase. In reality, the situation in real engines is more complex: the presence of a turbocharging system, which is usually overdesigned in most part of engine operation, ‘‘separates’’ engine exhaust immediately after the exhaust valves from engine exhaust downstream the turbine. Modifying back pressure at turbine exhaust and leaving boost pressure unchanged (engine load set by the ECU which operates on boost pressure), would require a new equilibrium of the turbo-compressor. This behaves differently if a waste gate system is used to control turbine power or if an IGV is operated. For this reason, a stronger scientific interest is needed on this recovery technology focusing the attention on the important interactions produced on the engine and on the vehicle. The Authors already developed an ORC-based power unit which is under an extensive testing [41]. It has a unique peculiarity related to the very low high temperature source (80–90 °C) to widen the potential applications and a novel technology for the expander based on rotary vane sliding machines; rated mechanical power recovered is close to 2 kW. In this paper, an IVECO F1C Turbocharged Diesel Engine 3.0 has been tested on a dynamometer test bench with the presence at the exhaust (after the catalyzer) of a Heat Recovery Vapor Generator (HRVG) which fed an ORC-based recovery unit. Main features of the power unit was the use of plate heat exchangers at the HRVG and at the condenser side and the use of rotary vanes machines both for the pump and the expander, as a result of a great research on these machines [42,43]. Sliding vane rotary machines, in fact, have same intrinsic advantages and they should be reconsidered in the panorama of technological choices: they are noiseless, compact, flexible from a geometrical point of view (diameter/length ratio), very reliable and do not require important maintenance actions [44,45]. They rotate at conventional revolution speeds (1500– 3000 RPM), suitable for electricity generation and an eventual speed increase produces a proportional size reduction and weight. The engine is fully equipped in order to measure engine properties and related relevant components (turbo-compressor speed, rack position, etc.). In fact, the engine presents a variable geometry turbine and its turbocharging group controlled by a variable vane (IGV) in order to adjust the pressure in the exhaust manifold [46,47]: this actuation is much more effective on the turbocharging group with respect to the by-pass of a part of the exhaust gas through a waste gate. At low gases flow rates (engine speed), the closure of the inlet guide vane speed up the gases themselves, increasing the reaction produced on the turbine wheel. In the tested engine, steady working conditions are considered in terms of torque at given revolution speeds. When a high back pressure occurs at turbine outlet (due to the plate type HRVG), the VGT actuator changes the position of the IGV and a new turbocharging speed is reached. Therefore, the boost pressure on its turn is modified and, definitively, engine characteristics in terms of fuel consumption, emission levels, etc. Moreover, the effects of the extra weight of the recovery unit was considered, when the engine considered is used to propel a light duty vehicle, in term of Brake Specific Fuel Consumption (BSFC) increase.
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In the paper; furthermore, the off design of the HRVG was studied occurring when the engine runs at a different steady point. A 1D model was developed in order to predict the thermodynamic conditions of the HRVG at the gas and working fluid sides, so evaluating the inlet conditions of the expander which is operated at off design conditions too. This knowledge was useful to preliminarily evaluating the role of fluid flow rate and its capability to insure heat recovery. From a previous experimental activity on an ORC-based power unit [21], which had the same inlet conditions at the expander as those estimated by the HRVG’s off design, the mechanical power recovered by the expander has been assessed. In this way, the net effect on the fuel consumption was discussed, contributing to redefine the real expectations of this technology.
2. The experimental apparatus The experimental campaign was performed on a dynamic engine test-bench system which consists of a combustion engine directly coupled with an electric dynamometer (AVL APA 100) and controlled by a real-time system (PUMA 5.6). The engine tested is the IVECO F1C 3.0 L. It is a turbocharged direct-injection diesel engine equipped to comply with the regulatory constraints EURO IV. The engine is equipped with a short route EGR system and its turbocharging group includes a variable geometry turbine (VGT), a stator stage that allows to modify the turbocharging speed in order to fulfill the boost pressure required by the engine. The electronic control unit (ECU) is a Bosch ETK P7 type. The detailed characteristics of the engine IVECO F1C are listed in Table 1. The engine was equipped with a series of sensors adapted to detect relevant engine variables, in order to monitor the behavior of the engine during the experimental tests. Thermocouples, thermistors and pressure sensors are installed in all the characteristic points of the engine intake and exhaust lines. A gravimetric balance is used to measure the fuel consumption and an air flow meter BOSCH EH-10523; the characteristic parameters of the turbocharging (boost pressure vs. engine speed and load), EGR and injection can be read from the ECU. The turbocharging compressor was equipped with an AVL Turbo-Speed Sensor TS350 in order to measure the speed of the compressor turbine. The measurements of the inlet and outlet pressure of the compressor and of the turbine allow the definition of the equilibrium point of the group. The aim of this project is to study the effect of the HRVG mounted on the exhaust line on the engine performances. The link between the ORC-based plant and combustion engine is represented by the back pressure produced by the HRVG mounted just after the catalyst (Fig. 2) in order to preserve the operating temperature of the catalyst itself. The ORC-based plant is also composed of a sliding vane rotary expander (connected to an electric generator), a pump, a water cooled condenser and a regenerator/economizer used to improve ORC efficiency (Fig. 2). Working fluid used is R236fa operating in the range of pressure of 3–12 bar.
Table 1 Specifications of the diesel engine used (IVECO F1C). Displaced volume Stroke Bore Connecting rod Compression ratio Number of valves Number of cylinders Maximum power Maximum torque
2998 cc 104 mm 95.8 mm 255 mm 19:1 16 4 in line 130 kW @ 3250 RPM 400 Nm @ 2000 RPM
All the heat exchangers (and, in particular, HRVG) are of a plate heat exchangers type. They present a good compactness, the ability to be modulated (simply adding one or more plates), ease of maintenance, high exchange efficiency, simplicity of assembly and transport, good resistance to high temperatures and corrosion, suitable shaping in order to simplify the on board installation. Fig. 3 shows the performances of the HRVG chosen in terms of pressure losses when it is crossed by a gas: pressure losses increases and reach 350 mbar when the gas flow rate reached about 500 kg/h. Overall dimensions are 243 393 130 mm with a volume of about 2.7 L. The high value of back pressure (more than 250 mbar) invites to a choice of an evaporative heat exchanger with greater permeability. In particular, it is noticeable the quadratic trend of the back pressure with gas flow rate which is closely linked to the speed rotation of the engine. The quadratic law (Fig. 3), can be expressed as:
Dp ¼ n kp m2gas
ð1Þ
where kp = 1.316 104 kPa/(kg/h)2 represents the best fit of the experimental data of Fig. 3 (n = 1) and n is an adimensional size parameter which allows to shape differently the pressure vs. flow rate curve. In fact, n = 1 represents the tested plate type HRVG assumed as a reference case. A value of n < 1 represents a family of more permeable (lower pressure losses for the same gas flow rate) HRVG designs which is suitable. This parametric dependency given by n will be useful for discerning the fuel consumption increase as a function of HRVG back pressure (and type) on the same engine. At design conditions, the organic fluid considered (R236fa) has a pressure of 12 bar, entering in the HRVG at 45 °C and exiting at 111 °C after a superheating. Its mass flow rate is about 131 g/s. At the same time, the engine operating point considered has a flow of gas equal to 0.072 kg/s and a temperature of the gases at the catalyst of 330 °C, this corresponds to an engine operating condition equal to 100 Nm @ 2500 RPM. This engine working point, being a medium–low one, is a characteristic of a homologation driving cycle of a vehicle (passenger car or light duty). In order to keep the heat exchanger dimensions under control, a suitable pressure loss has been considered about 60 mbar at design conditions at the working fluid side. The heat exchangers are sized according to the method of Logarithmic Mean Difference Temperatures which permitted to determine the heat exchange surface of the heat exchanger. The heat exchanger was produced by the company EmmeGi S.p.A. with the characteristics shown in Table 2. As already observed, the Variable Geometry Turbocharger (VGT) raises the boost pressure even at lower engine speeds, together with the reduction of engine pumping losses at higher engine speeds, compared with a waste-gated turbocharger. VGT is a device that can vary the flow area and flow angle between the turbine volute and rotor channel. At low engine speeds, making the flow passage between turbine nozzle vanes narrower, the exhaust gas approaching the turbine rotor channel is accelerated. By this acceleration, the boost pressure and, therefore, the charge air flow rate is increased. At high engine speeds, the flow rate of the gas is higher and it is not necessary to speed up the flow to move the compressor; therefore, the area of the flow passage between nozzle vanes can be larger and it is realized by manipulating the IGV angle adequately. This results in the reduced pumping loss, which is the primary reason of the lower fuel consumption with VGT at high engine speeds. In a medium speed range, with raising the boost pressure and increasing the charge air mass, higher torque can also be obtained. But, the performance level in this region is already limited to the maximum cylinder pressure [48].
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Fig. 2. Engine scheme with heat recovery vapor generator on the exhaust line.
Table 3 Back pressure increase due to the HRVG on the exhaust line in the working points considered.
Fig. 3. Experimental curve of HRVG pressure drops vs. gas flow rate mgas.
Table 2 Characteristics of the heat exchanger tested.
Number of channels Capacity Overall heat exchange surface Heat transfer coefficient Plates Fouling Pinch point
Hot side
Cold side
25 16 kW 4.46 m2 456/117 W/m2 °C 50 6.19 m2 °C/kW 55 °C
24
3. Engine experimental campaign and results The experimental tests on IVECO F1C engine were conducted in steady conditions reproducing the engine characteristic curve at
Engine torque (Nm)
Engine speed (RPM)
Exhaust mass flow rate (kg/h)
Exhaust gases temperature after catalyst (°C)
Exhaust back pressure (after catalyst) (mbar)
112 127 187 191 198 202 199 202 200 198 186 173 153 110
835 1000 1250 1500 1800 2000 2250 2500 2750 3000 3250 3500 3750 3900
76.3 91.7 133.3 166.8 239.8 275.1 310.1 336.6 369.8 384.7 417.5 455.5 512.7 454.5
243.1 275.6 336.5 341.7 337.1 340.3 360.1 371.8 373.7 415.7 439.9 460.7 430.8 400.2
12 25 37 52 90 107 125 168 188 209 235 279 335 266
50% of maximum load for different revolution speed: Table 3 shows all the engine point tested and, in particular, exhaust mass flow rates and temperatures of the gas at the inlet of the HRVG. First, tests were carried out in OEM condition and, then, they were conducted again with the HRVG placed after the catalyst in the exhaust line. Focus has been given at the thermodynamic conditions measured in this point. Temperature has been measured by K-Thermocouple and pressure by membrane transducers. Table 3 summarizes the increase of the pressure measured in this point of the exhaust line due to the presence of the heat exchanger: it reaches values above 300 mbar. In particular, this pressure increase is correlated to the gas flow rate that crosses the exhaust pipeline, which depends on the engine speed and intake manifold air pressure. Table 3 shows how the back pressure rises according to the engine speed, except to the last operating point considered, where the lower torque is related to a lower charge air pressure and, consequently, to a lower air mass flow rate. Therefore, pressure increase is reduced, even though engine speed increases. It is expected that this back pressure increase would imply an extra fuel consumption, in order to keep the mechanical power
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of the engine. An average value of fuel consumption increase is demonstrated of about 2% (3.95 g/kW h), but it can be very high (till to 5%) at higher engine speeds. The propensity for the BSFC to increase (Fig. 4), which was observed at all operating conditions, was attributed to ineffective blow-down of the engine exhaust, causing more residual gas that remains in the cylinders [49]. Consequently, the charging efficiency decreases, and the engine must consume more fuel to give the same mechanical power. Data on Fig. 3 and on Fig. 4 can be analytically processed in order to correlate the BSFC with the gas flow rate, thus, ultimately, with the pressure losses across the HRVG (back pressure). Therefore, different families of HRVG represented by different values of the size parameter n can be considered and studied in terms of the BSFC increase that they would produce on the tested engine. Fig. 5 reports the exhaust flow rates variations as a function of the engine speed. Best fit of these variations is given by a straight line (Eq. (2)). Exhaust back pressure does not significantly affect the mass flow rate.
mgas ¼ km N
ð2Þ
In this way, the dependency of the BSFC vs. back pressure can be analytically expressed as in Eq. (3), and, so, related to the permeability parameter n (Eq. (1)).
DBSFC ¼ kc;1 Dp2 þ kc;2 Dp þ kc;3
ð3Þ
Table 4 specifies the parameters in Eqs. (3) and (4). n parameter reveals itself very useful because it simply reshapes pressure losses and it represents a more permeable HRVG, Fig. 6. Taking as reference the tested plate-type HRVG (for which n = 1, Eq. (1)), the n < 1 range reports the effect of a more permeable HRVG on the engine BSFC due to the back pressure produced. The n > 1 range represents the behavior of a less permeable HRVG (greater pressure losses produced with respect to the actual design). It is evident that for engine speeds less than 2000 RPM, the effect of the back pressure on the BSFC is below 1.5%. For engine speeds greater than this value, the back pressure effect starts to be great and makes nil the benefits of the energy recovered by the ORC-based unit. Fig. 6, therefore, can be used to evaluate the negative effect produced by the back pressure on the engine tested and should be taken into consideration to evaluate the net effect of the energy recovery unit ORC based. The back pressure produced by the HRVG rises back through all the exhaust line from the catalyst till to the turbine (Fig. 2). This back pressure increase, however, seems to affect to a lesser extent the turbine inlet (Fig. 7): relative pressure differences between the two cases are very low till to 1800 RPM, only at high speed point the pressure increases significantly. This signifies that, in some way, the VGT is able to rearrange its working point and the pressure increase in the engine exhaust manifold is not as high as the back pressure after the turbine. The VGT mitigates the fuel consumption increase due to the back pressure, but it cannot completely compensate this negative effect. Therefore, the repositioning of the IGV at the inlet of the turbine (just after the exhaust manifold) defines a new equilibrium point of the engine which corresponds to a different boost pressure of the air entering in the cylinders. Fig. 8 illustrates the boost pressure increase due to the HRVG inside the intake common manifold for the operating conditions analyzed. This behavior is, however, so complex: at low engine speeds the boost pressure is quite the same in both cases (with and without
Table 4 Calibration parameters of Eqs. (2) and (3).
Fig. 4. BSFC increase in presence of the heat exchanger for the operating conditions (50% of the maximum engine load).
Fig. 5. Exhaust mass flow rates vs. engine speed: experimental values with and without HRVG compared with a linear variation.
km kc,1 kc,2 kc,3
0.1183 4.993 105 4.687 103 1.1005
kg/h/RPM g/kW h/mbar2 g/kW h/mbar g/kW h
Fig. 6. Computed BSFC of a different permeable HRVG based on measured value on the IVECO F1 engine.
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Fig. 7. Comparison between exhaust manifold pressure in the two cases for the operating conditions examined (50% of the maximum engine load).
Fig. 8. Boost pressure in the two cases for the operating conditions examined (50% of the maximum engine load).
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Fig. 9. Comparison between turbocharger revolution speeds difference and rack position difference between the two cases (positive rack position differences indicates a more open VGT, while negative ones a more closed VGT).
related to the component themselves (expander, pump, heat exchangers, piping, fluid inside, etc.) but also to the additional radiator which is needed in order to deliver outside the heat removed at the condenser. The recovery unit, in fact, re-introduces inside the ‘‘engine system’’ the heat that would have been reversed toward the environment. Therefore, the estimation of the weight increase depends on many factors; among them, the size of the power unit and the final destination of the heat removed from the condenser. Fig. 10 shows the vehicle propulsive power increase due to the additional weight of a ORC-based plant placed on board, for the typical data of a light duty vehicle (Table 5). Propulsive power can be easily estimated considering the equilibrium of the forces on a vehicle during motion in steady conditions. Vehicle driving conditions are calculated from the engine working point of Table 3 (engine torque and speed), considering the proper gear ratio of the vehicle considered. The influence of the weight increase is much higher at low engine speeds, which are definitely the most used in real driving cycles.
HRVG presence), while at higher engine speeds the system turbine-compressor increases the boost pressure of the engine, restoring the air mass inside the cylinder when the engine back pressure tends to grow up. This is confirmed by the value of the position of the VGT actuator (‘‘rack’’): till to 1800 RPM the VGT is fully closed and it is not influenced by the presence of the HRVG (Fig. 9); therefore, the higher back pressure, due the HRVG, leads to a lower turbine specific power, slowing the turbocharger group (Fig. 9). On the other hand, at higher engine speeds, the VGT is more closed, when the HRVG is present, and turbocharger is accelerated (Fig. 9), producing a significant boost pressure increase (Fig. 8). Definitely, when a plate heat exchanger (characterized by the pressure losses as in Fig. 3) is present on the exhaust, engine boost pressure increases till to 100 mbar (depending on the engine speed), in order to restore the engine torque, turbocharger speed increases in the most part of the engine speed range even though, due to the IGV rack position, till to 1800 RPM it slightly decreases. 3.1. Weight increase An additional important issue related to the ORC-based power unit presence on a vehicle is its not-negligible weight. This is
Fig. 10. Propulsive power increase due to the additional ORC-plant weight on the vehicle for a light duty vehicle (engine torque is 50% of the maximum one). Data on the legend are referred to propulsive power considered for each curve.
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can be evaluated and how the temperature difference DT between them proceeds inside the heat exchanger. Therefore, the overall heat exchanged Qexch can be evaluated. Calculation is completed by pressure drops: at the gas side they have been experimentally evaluated, while at the fluid side they have been calculated according to homogeneous flow Friedel model [51]. Hence, the deviations with respect to the pressure drop at design conditions have been evaluated. Outlet quality x of the refrigerant is defined as in Eq. (5):
Table 5 Light-duty vehicle reference data. Vehicle weight Drag coefficient Frontal area Tires radius Tires friction coefficient
3350 kg 0.5 4.3 m2 0.372 m 0.017
For an additional weight of 50 kg, for instance, the propulsive power increase is of about 0.7% at 3500 RPM and reaches 1.25% at 1000 RPM. This is paid in terms of fuel consumption and emissions increase.
x¼
Hout Hsl Hsv Hsl
ð5Þ
where subscript sl is saturated liquid and sv is saturated vapor. Therefore, negative quality means subcooled fluid, while x > 1 means a superheated vapor. The HRVG off design was calculated using:
3.2. Heat recovery vapor generator off-design analysis A correct evaluation of the power recoverable from the gases must consider the off design behavior of the recovery unit, mainly of the HRVG and of the expander. This is because of the strong flow rate and temperature variations of the gases during engine operation. In order to take into account these principal effects (gas flow rate and gas temperature), a 1D model of the HVRG was considered making reference to a plate heat exchanger properly sized: The fluid path has an evident 1D nature and it was divided into elementary pieces in which mass, momentum and energy conservation equations in steady conditions were simultaneously solved (Eq. (4)). This was done both at the gas and at the working fluid sides, considering a counterflow heat exchanger. The cross section of the fluids passages were known being the heat exchanger already sized for a specific design flow rate and temperature inlet (at the gas side and at the working fluid side).
8 m ¼ qin Ain v in ¼ qout Aout v out > > > > v 2in > v 2out > > < m 2 þ mHin þ Q exch ¼ m 2 þ mHout mv in þ pin Ain ¼ mv out þ pout Aout > > > > pout pin ¼ f ðm; qin ; qout Þ > > > : þ hfluid Þ S DT Q exch ¼ ðhgas þ kwall t
(a) From the gas side, the results of the experimental characterization of the engine in terms of flow rate and temperature (fourteen points, Table 6); the design conditions were 72 g/s and 330 °C as flow rate and inlet temperature. (b) From the refrigerant side the design conditions were 45 °C and 131 g/s as temperature inlet and flow rate, respectively. They have been considered constant during the off design conditions at the gas side; evaporating pressure is about 12 bar. Table 6 summarizes the HRVG off-design behavior when the refrigerant mass flow rate is equal to 131 g/s. In particular, the outlet quality of the R236fa gives the status of vaporization of the organic fluid. At the lowest conditions (tests #1 and #2) in terms of gas flow rate and temperature, the vaporization does not take place: only a part of the economization of the fluid is reached (without reaching the saturation liquid state). When flow rates increase and temperature too (tests #3 and #4), a liquid–vapor saturated mixture is produced while for a further increase (tests #6 to #10 and #14) a good range of superheating is reached. For an additional increase (tests #11 to #13) the superheating becomes too high, incompatible with the degradation of the organic fluid (temperature limit is close to 230 °C). In particular, Fig. 11 shows the thermal exchange diagram of the HRVG close to the design conditions (test #6): the organic fluid is completely evaporated and reaches a superheating of 10 °C. In Fig. 12 a low thermal power exchanged case is represented (test #4): the evaporation is just started when the organic fluid exits from the HRVG. Fig. 13 shows a case where the power exchanged by the gas is very high: the result is that the organic fluid reaches a great value of superheating, very close to the degradation point of the fluid.
ð4Þ
Starting from the inlet thermo-fluid-dynamic conditions (pressure pin, density qin, enthalpy Hin, temperature Tin, mass flow rate m) and knowing the heat exchanger geometry and material used (cross sections Ain and Aout, thickness t, thermal conductivity kwall, heat exchange surface S), the convective heat transfer coefficients hgas and hfluid can be calculated for each HVRG elementary piece according to explicit correlations: Dittus–Boelter correlations are used in single phase fluids, while Kandlikar specific plate correlation is used for evaporation [50]. Following in counterflow the two fluids, for each elementary piece, the temperature of the fluids Table 6 HRVG operating conditions in all the test effectuated. Test # Exhaust mass flow rate Exhaust pressure drop Exhaust inlet temperature Exhaust outlet temperature Exhaust thermal power exchanged R236fa mass flow rate R236fa inlet temperature R236fa outlet temperature R236fa outlet quality R236fa outlet density Heat exchanger effectiveness
g/s bar °C °C kW g/s °C °C Fraction kg/m3 %
1
2
3
4
5
6
7
8
9
10
11
12
13
14
21.2 0.283 242.9 46.3 4.6 131.2 45 67.4 0.15 1202 96.7
25.5 0.284 268.0 46.1 6.3 131.2 45 74.3 0.06 1168 93.5
37.0 0.255 348.0 51.0 12.3 131.3 45 78.7 0.40 194.5 98.8
46.3 0.256 342.5 51.4 15.1 131.3 45 78.7 0.55 146.8 97.1
66.6 0.265 346.0 54.0 21.9 131.3 45 78.7 0.94 91.6 95.0
76.4 0.258 358.0 64.5 25.4 131.3 45 111.2 1.34 70.0 99.4
86.1 0.267 370.0 69.3 29.5 131.3 45 140.0 1.63 61.6 99.4
93.5 0.286 383.0 72.9 33.2 131.3 45 158.8 1.84 57.5 100
102.7 0.289 385.0 77.6 36.3 131.3 45 185.5 2.10 52.8 99.1
106.9 0.314 436.0 83.2 43.7 131.2 45 229.5 2.58 46.8 98.0
116.0 0.348 456.8 91.4 49.4 131.3 45 232.2 2.61 46.4 98.6
126.5 0.362 470.1 101.0 54.8 131.2 45 232.2 2.61 46.4 98.5
142.4 0.379 423.0 102.3 53.5 131.2 45 232.2 2.61 46.4 97.7
126.3 0.325 389.0 88.6 44.1 131.2 45 230.8 2.59 46.6 97.5
D. Di Battista et al. / Applied Energy 152 (2015) 109–120
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Fig. 11. Thermal exchange curves in design conditions (test #6). Fig. 13. Thermal exchange curves in high power conditions (test #14).
which is of a rotary volumetric type so linearly correlated to the pump speed. Table 7 summarizes the results with these three different flow rates: fluid degradation is always avoided and working fluid at the outlet of the evaporator is almost always superheated, so suitable for expansions. Only for the tests #1 and #2 (at very low engine speed) refrigerant quality is too low for a good expansion: an even lower flow rate should be considered but the thermal power recovered by the gas would be really low. Finally, results obtained demonstrate that a model based control strategy applied to the working fluid flow rate is indeed suitable to manage the recovery unit. 3.3. Expander off-design analysis Also the expander is strongly affected by a variable behavior at off-design conditions. In particular, if it is entrained at fixed speed imposed by an electrical generator, the mass flow rate of the working fluid mWF has to change according to the expander inlet density qWF but this could not match with the HRVG requirements. For the application considered, working fluid flow rate is known (imposed) and the expander revolution speed has to change according to: Fig. 12. Thermal exchange curves in low power conditions (test #4).
Nexp ¼ It is evident how the designed refrigerant mass flow rate (131 g/s) is suitable only at medium–low load and speed conditions. Data in Table 6 highlights three ranges of working conditions among the fourteen points considered: tests from #1 to #4 have a low thermal power exchanged by the exhaust gases and working fluid flow rate is unsuitable, tests from #4 to #8 have a good behavior with the designed fluid mass flow rate, and tests from #9 to #14 which have a very high thermal power exchanged would require a greater working fluid flow rate. Hence, three suitable values of the refrigerant mass flow rate have been detected in order to introduce the need of controlling the working fluid flow rate: for the first range of working points (#1 to #4) it was reduced by 50%, for the second range (#4 to #8) it was kept as it was designed and, finally, for the third range (#9 to #14) it was increased by 50%. This flow rate, on the other hand, can be easily changed because it is imposed by the pump
mWF 60
qWF V suc nv anes
ð6Þ
for an ideal vane filling. Inlet fluid density remains the only variable to match. Fig. 14 shows how expander speed has to change in order to match inlet density, at different working fluid flow rates. Fig. 14 allows to observe that expander speed should change sensibly in order to match the requirements of the HRVG and those of a perfect filling of the expander. At low thermal engine power recovered, expander speed should be in the range of 250– 1500 RPM while at rated design conditions it increases to 2000– 3500 RPM. For a higher thermal power recovered, best expander speed is in the range of 4000–6000 RPM. Dynamic expanders appear not immediately suitable to allow this requirement while volumetric machines are more flexible. Among them, rotary vanes introduce further additional advantages in terms of fluid management, robustness, noise, shape factor sizing, reliability and ease of maintenance.
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Table 7 HRVG operating conditions in all the test effectuated with a three suitable refrigerant mass flow rate (low, design value, high). Test # Exhaust mass flow rate Exhaust pressure drop Exhaust inlet temperature Exhaust outlet temperature Exhaust thermal power exchanged R236fa mass flow rate R236fa inlet temperature R236fa outlet temperature R236fa outlet quality R236fa outlet density Heat exchanger effectiveness
g/s bar °C °C kW g/s °C °C Fraction kg/m3 %
1
2
3
4
5
6
7
8
9
10
11
12
13
14
21.2 0.015 242.9 47.4 4.6 65.6 45 77.7 0.15 409.9 95.4
25.5 0.018 268.0 46.8 6.2 65.6 45 78.7 0.36 210.7 95.8
37.0 0.034 348.0 50.0 12.2 65.6 45 78.7 1.11 79.7 94.6
46.3 0.048 342.5 52.5 14.8 65.6 45 78.7 1.39 68.1 93.3
66.6 0.085 346.0 54.0 21.9 131.3 45 78.7 0.94 91.6 95.0
76.4 0.105 358.0 64.5 25.4 131.3 45 111.2 1.34 70.0 99.4
86.1 0.127 370.0 69.3 29.5 131.3 45 140.0 1.63 61.6 99.4
93.5 0.145 383.0 72.9 33.2 131.3 45 158.8 1.84 57.5 100
102.7 0.169 385.0 67.7 37.6 196.9 45 124.6 1.33 70.3 99.7
106.9 0.182 436.0 72.7 45.2 196.9 45 163.5 1.71 59.8 100.0
116.0 0.216 456.8 77.8 51.6 196.9 45 194.3 2.01 54.2 99.9
126.5 0.262 470.1 83.8 57.8 196.9 45 222.6 2.29 50.1 99.7
142.4 0.393 423.0 85.7 56.6 196.9 45 216.9 2.23 50.9 99.7
126.3 0.261 389.0 76.0 46.1 196.9 45 167.5 1.74 59.2 99.8
Fig. 15. BSFC variations due to back pressure increase, 50 kg weight increase on board vehicle and mechanical power recovered by ORC-based plant. Fig. 14. Expander speed ad a function of the fluid density at the expander inlet (expander vanes are served and suction volume is Vsuc is 5.9 cm3).
Gas flow rate by pass as well as other strategies which consider a multivariable control can match the HRVG and the expander constraints. Data from HRVG off design conditions in Table 6 were used to evaluate the performances of an ORC-based power unit previously developed by the Authors [21]. Among the wide available experimental data set, those which match closer the conditions presented in Table 6 regarding flow rates and inlet thermodynamic conditions at the expander have been considered. This procedure appears to be consistent because the performances of the power unit are the same, regardless of the high temperature source, if flow rate and inlet expander conditions are kept equal. In this way the mechanical power really recoverable could be estimated as the HRVG was fed with the exhaust gases of the F1C IVECO engine. Finally, Fig. 15 shows the compared effects on the BSFC of the back pressure, the effect of an extraweight of 50 kg of the ORC plant and the mechanical power recovered by the expander. The first produced an increase of the BSFC (Fig. 4) as for the second, while the third has a positive effect: the overall mechanical power is the sum of the engine shaft power and the ORC mechanical power recovered. Mechanical power recovered by ORC-plant is evaluated considering the exhaust thermal energy Qexhaust and experimental overall ORC efficiency, considering the expander in off-design conditions [21] (Eq. (7)).
PORC ¼ gORC Q exhaust
ð7Þ
In Fig. 15 is evident how the BSFC reduction due to the mechanical power recovered by the ORC plant is always higher than the BSFC increase due to back pressure effect on the engine. The net
positive effect depends on the engine speed which reports the effect of the thermal power recovered. Data represented in Fig. 15 have been reported in Table 8 in terms of BSFC percentage increase (or decrease), vs. engine speed. BSFC reference values are those of the original engine equipment. From values reported in Table 8 it is clear that the weight of the ORC determines an increase of the fuel consumption because of the propulsion power increase. It tends to decrease with the increase of the engine speed: Table 8 shows a value close to 1% but a more precise estimation depends on the weight increase produced by the power unit, with respect to the original vehicle weight. The effect of the back pressure produced on the exhaust line of the engine causes a variable trend of the BSFC increase and it is strongly engine dependent, as discussed previously. Average increase value for the testing activity done is close to 3.6% but this is related to the permeability of the plate heat exchanger tested. Other HRVG types could produce lower values: a reference ideal situation can be estimated considering nil this contribution. Table 8 demonstrates definitively that a big difference exists between potential recovery evaluated by only thermodynamic considerations and the real one, which accounts for the present Table 8 BSFC percentage increase/decrease vs. engine speed. RPM
Back pressure (%)
Weight increase (%)
Mechanical recovery (%)
Net (%)
1000 1500 2000 2500 3000 3500
0.19 0.60 1.46 0.89 0.90 2.82
1.20 1.06 0.92 0.87 0.78 0.72
1.56 2.02 2.70 3.00 4.08 6.32
0.17 0.37 0.31 1.24 2.40 2.78
D. Di Battista et al. / Applied Energy 152 (2015) 109–120
technology available and for the interferences produced on the engine: the poor values of the recovery are also associated to the strong off design suffered by the ORC-based power unit whose initial design point must be chosen with care. Considering a negligible contribution due to the back pressure, a maximum net benefit is close to 5.5%. Finally, an additional loss must be considered represented by the energy required to drive the fan which would guarantee the required heat exchange toward the environment, if needed.
4. Conclusions The heat recovery from the exhaust gases in ICE has an interesting potential and it is characterized by a cost increase per gram of CO2 saved very promising with respect to the other technologies. When an ORC-based power unit is considered, four contributions need to be evaluated to state the net recovery’s benefit: (a) the effect of the back pressure produced on the engine by the HRVG; (b) the weight increase of the vehicle due to the power unit; (c) the power recovered by the unit taking into consideration the off design conditions of the HRVG and that of the expander; (d) the additional power required to discharge the energy extracted at the condenser toward the environment. The final use of the net power recovered as in electrical form or in mechanical form is an additional issue. Making reference to an IVECO F1C 3.0 L turbocharged diesel engine operated in an engine test bed, the paper discussed: (a) The effect of the back pressure produced by a real HRVG which fed an ORC-based power unit. The presence of an IGV turbine control makes the prediction of these effects not straightforward. (b) The effect produced by the weight increase due to the ORC-based power unit. (c) The recovery done by an existing power unit taking into account the off design situations produced on the HRVG and on the expander by engine speed variation. The engine was fully sensorized in a way to understand how it reacts to a back pressure, mainly in terms of the turbocharging system with and IGV control. Main conclusions are: (a) Engine back pressure has a crucial role and can represent the most sensible interference of the heat recovery system on the engine. Plate heat exchangers are not suitable for this application: more permeable components are suggested, i.e. shell and finned tube. Limiting the back pressure at 175 mbar for a flow rate equal to 500 kg/h, a negligible effect is done on the BSFC (less than 1% in the worst case), thanks also to the IGV system. (b) When a plate heat exchanger is characterized by a pressure loss equal to 350 mbar at 500 kg/h (similar to the one tested), engine boost pressure increases till to 100 mbar depending on the engine speed. This is required by the engine in order to restore previous engine torque, turbocharger speed increases in the most part of the engine speed range even though, due to the IGV rack position, till to 1500 RPM it slightly decreases. (c) The tested counterflow HRVG, even though properly designed. produces an unacceptable specific fuel consumption increase in the range of 2–5%. Greater values corresponds to higher gas flow rates. (d) Considering other heat exchanger types (for instance, shell and finned tubes) the BSFC increase can be kept below 1% at all engine speeds (i.e. engine flow rates).
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(e) The variation of the exhaust gas flow rate and inlet temperature at the inlet of the HRVG produces strong off-design conditions which must be managed with care. If the working fluid flow rate remains constant (at a suitable design value), vaporization could not occur as well as maximum allowable temperature could be reached implying fluid stability problems. (f) The HRVG off design conditions at the working fluid side entrain the expander to a similar off design which influences volumetric efficiency, reducing the mechanical power recoverable. The effects of both HRVG and expander off-design can be minimized thanks to a multivariable control in which the working fluid density, the mass flow rate, the expander speed as well as the exhaust gas flow rate can be changed. (g) 50 kg as extra-weight compared to the original equipment vehicle, whose weight was 3350 kg (light duty vehicle) appears correctly evaluated standing the present technology: this issue produces an additional power request which is equivalent to a loss of about 1% in terms of fuel consumption. Extra weight assumed considers also the bigger radiator needed to reverse toward the environment the heat exchanged at the condenser. (h) Gross benefit of the ORC-based unit power is of level of 4%– 5%; additional improvement is possible by an appropriate choice of the design point of the unit and model based control. A final consideration which further limits the net recovery is related to the need to discharge toward the environment the heat removed at the condenser. The presence of an additional fan must be taken into account and when it is operated (low vehicle speed and high external temperature) an additional loss has to be considered. A further limitation is related to the condenser temperature which depends, definitely, on external temperature. A potential improvement is under development concerning: (a) the use of a more permeable HRVG; (b) the downsizing of the power unit and its weight reduction: (c) the design of more efficient expander; (d) the reduction of the off design effects through the choice of a more suitable design point and through a model based control of the recovery unit. Acknowledgements Meccanotecnica Umbra S.p.A. and Ing. Enea Mattei S.p.A. are acknowledged for continuous technological support and funding. References [1] EPA. Light-duty automotive technology. Carbon dioxide emissions, and fuel economy trends: 1975 through 2014 – October 2014. [2] Chiara F, Canova M. A review of energy consumption, management and recovery in automotive systems with considerations on future trends. Proc Inst Mech Eng Part D: J Automob Eng 2013;227(6):914–36. [3] Zhang Y, Zhao H. Investigation of combustion, performance and emission characteristics of 2-stroke and 4 stroke spark ignition and CAI/HCCI operations in a DI gasoline. Appl Energy 2014;130:244–55. [4] Conklin James C, Szybist James P. A highly efficient six-stroke internal combustion engine cycle with water injection for in-cylinder exhaust heat recovery. Energy 2010;35(4):1658–64. ISSN 0360-5442. [5] Cipollone R, Di Battista D, Gualtieri A. A novel engine cooling system with two circuits operating at different temperatures. Energy Convers Manage 2013;75:581–92. ISSN 0196-8904. [6] Cipollone R, Di Battista D. Performances and opportunities of an engine cooling system with a double circuit at two temperature levels. SAE technical paper 2012-01-0638; 2012 http://dx.doi.org/10.4271/2012-01-0638. [7] Saidur R, Rezaei M, Muzammil WK, Hassan MH, et al. Technologies to recover exhaust heat from internal combustion engines. Renew Sustain Energy Rev 2012;16(8):5649–59. ISSN 1364-0321. [8] Singh S, Garg A, Gupta A, Permude A. Analysis of thermal balance of diesel engine and identification of scope for waste heat recovery. SAE technical paper 2013-01-2744; 2013 http://dx.doi.org/10.4271/2013-01-2744.
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