C Hahn

G De Loubens

O Klein

M Viret

V V Naletov

J Ben Youssef

Comparative Measurements of Inverse Spin Hall and Magnetoresistance in YIG|Pt and YIG|Ta

We report on a comparative study of spin Hall related effects and magnetoresistance in YIG|Pt and YIG|Ta bilayers. These combined measurements allow to estimate the characteristic transport parameters of both Pt and Ta layers juxtaposed to YIG: the spin mixing conductance G ↑↓ at the YIG|normal metal interface, the spin Hall angle ΘSH, and the spin diffusion length λ sd in the normal metal. The inverse spin Hall voltages generated in Pt and Ta by the pure spin current pumped from YIG excited at resonance confirm the opposite signs of spin Hall angles in these two materials. Moreover, from the dependence of the inverse spin Hall voltage on the Ta thickness, we extract the spin diffusion length in Ta, found to be λ Ta sd = 1.8 ± 0.7 nm. Both the YIG|Pt and YIG|Ta systems display a similar variation of resistance upon magnetic field orientation, which can be explained in the recently developed framework of spin Hall magnetoresistance.

I. INTRODUCTION

Spintronics aims at designing devices which capitalize on the interplay between the spin-and charge-degrees of freedom of the electron. In particular, it is of central interest to study the interconversion from a spin current, the motion of spin angular momentum, to a charge current and the transfer of spin angular momentum between the conduction electrons of a normal metal (NM) and the magnetization of a ferromagnetic material (FM). The separation of oppositely spin polarized electrons of a charge-current through spin-orbit-coupling is called spin Hall effect (SHE) 1,2 . Its inverse process (ISHE) converts spin-currents into charge-currents and has recently sparked an intense research activity 3,4 , as it allows for an electrical detection of the dynamical state of a ferromagnet 5,6 . Indeed, a precessing magnetization in a ferromagnet generates a spin current via spin pumping 7 , which can be converted, at the interface with an adjacent normal layer, to a dc voltage by ISHE. Moreover, electronic transport can also be affected by the static magnetization in the FM as electrons spins separated by SHE can undergo different spin-flip-scattering on the interface with the FM layer. In particular, spin flipped electrons are deflected by ISHE in a direction opposite to the initial current, leading to a reduced total current at constant voltage. This effect depends on the relative orientation between magnetization and current direction, and has recently been called spin Hall magnetoresistance (SMR) 8 .

Experimental studies on spin pumping induced inverse spin Hall voltages (V ISH ) in FM|NM bilayers were first carried out with Pt as NM in combination with NiFe as FM 5,[9][10][11][12] and more recently with the insulating ferrimag-net Yttrium Iron Garnet (YIG) 6,[13][14][15][16][17] . Although other strong spin-orbit metals have been tried in combination with the metallic ferromagnets NiFe 18,19 and CoFeB 20,21 , inverse spin Hall voltage 6,17 and magnetoresistance 8,22 measurements made on YIG|NM have so far been limited to NM=Pt. Still, it would be very interesting to compare V ISH and SMR measurements on different YIG|NM systems, including metals having opposite spin Hall angles, such as Pt vs. Ta 20,23 . Ab initio calculations indeed predict the spin Hall angle of the resistive β-phase of Ta to be larger and of opposite sign to that of Pt 24 . The defining parameters for V ISH and SMR are the spin diffusion length in the normal metal (λ sd ), the spin Hall angle (Θ SH ) which quantifies the efficiency of spin-to charge-current conversion, and the spin mixing conductance (G ↑↓ ) which depends on the scattering matrices for electrons at the FM|NM interface 7 and can be seen as the transparency of the interface for transfer of spin angular momentum 25 . Evaluation of the three above mentioned parameters is a delicate task 26 , as the measured V ISH voltages and SMR ratio depend on all of them.

In this paper, we present a comparative study of YIG|Pt and YIG|Ta bilayers, where we measure both the ISHE and SMR on each sample. We confirm the opposite signs of spin Hall angles in Pt and Ta and the origin of SMR, which has been explained in Ref. 8 . Thanks to these combined measurements, we can evaluate the spin mixing conductances of the YIG|Pt and YIG|Ta interfaces and the spin Hall angles in Pt and Ta. In order to get more insight on the previously unexplored YIG|Ta system, we study the dependence of ISHE on Ta film thickness, which enables us to extract the spin diffusion length in Ta.

The remaining of the manuscript is organized as fol- lows. Section II gives details on the samples and experimental setup used in this study. In section III, the experimental data of V ISH and SMR obtained on the YIG|Pt and YIG|Ta systems are presented and analyzed. In section IV, we discuss the transport parameters extracted from our measurements. We also comment on the absence of direct effect of a charge current in Pt on the linewidth of our 200 nm thick YIG samples. Finally, we emphasize the main results of this work in the conclusion.

II. EXPERIMENTAL DETAILS

A. Samples

YIG films

Two single crystal Y 3 Fe 5 O 12 (YIG) films of 200 nm thickness were grown by liquid phase epitaxy on (111) Gd 3 Ga 5 O 12 (GGG) substrates 27 , and labeled YIG1 and YIG2. Epitaxial growth of the YIG was verified by Xray diffraction and the films roughness was determined by atomic force microscopy to be below 5 Å. Their magnetic static properties were investigated by vibrating sample magnetometry. The in-plane behavior of the thin YIG films is isotropic with a coercitivity below 0.6 Oe 27 . The saturation magnetization, found to be 140 emu/cm 3 , corresponds to the one of bulk YIG. This value was verified by performing ferromagnetic resonance (FMR) at different excitation frequencies.

FMR also allows to extract the magnetic dynamic properties of the 200 nm thick YIG films. A typical FMR spectrum of the YIG1 film obtained at 10 GHz and low microwave power (P = -20 dBm) is presented in Fig. 1a. The gyromagnetic ratio of our YIG films is found to be γ = 1.79 • 10 7 rad/s/Oe. From the dependence of the linewidth on the excitation frequency, their Gilbert damping α G = (2.0 ± 0.2) • 10 -4 can be determined, see Fig. 1b. This value highlights the very small magnetic relaxation of these thin films. Still, there is an inhomogeneous part to the linewidth (∆H 0 = 0.4 Oe in Fig. 1b). For one of the two prepared films (YIG2), two to three closely spaced resonance lines could be observed in some cases, which we attribute to distinct sample regions having slightly different properties.

YIG|Pt and YIG|Ta bilayers

After these standard magnetic characterizations, the YIG films were cut into slabs with lateral dimensions of 1.1 mm × 7 mm in order to perform inverse spin Hall voltage and magnetoresistance measurements. Platinum and tantalum thin films were then grown by sputter deposition, at a power density of 4 W/cm 2 . The growth of the resistive β-phase Ta was achieved by optimizing the Ar-pressure during the sputtering process. The appearance of this tetragonal crystalline phase in a narrow window around 10 -2 mbar was verified by the presence of characteristic lines in the X-ray diffraction spectra. The β-phase was also confirmed by the resistivity of the films 20 , which for 10 nm Ta thickness lies at 200 µΩ•cm.

In order to compare ISHE and SMR on YIG|Pt and YIG|Ta bilayers, a 15 nm thick Pt and a 3 nm thick Ta layers were grown on the YIG1 sample. The conductivities of these metallic films are σ Pt = 2.45 • 10 6 Ω -1 •m -1 (in agreement with the values reported in Refs. 17,18 ) and σ Ta = 3.05 • 10 5 Ω -1 •m -1 , respectively. These two samples have been used to obtain the results presented in Figs. 2 and4. The dependence on Pt thickness of both V ISH 17 and magnetoresistance 22,28 has been studied earlier. In this work, we have used the YIG2 sample to study the dependence as a function of the Ta thickness, which was varied from 1.5 nm to 15 nm (1.5, 2, 3, 5, 10 and 15 nm). The conductivity of these Ta films increases from 0.8 • 10 5 Ω -1 •m -1 to 7.5 • 10 5 Ω -1 •m -1 with the film thickness. This series of samples has been used to obtain the data of Fig. 3. Finally, Pt films with thicknesses 10 nm and 15 nm were also grown on YIG2, for the sake of comparison with YIG1.

B. Measurement setup

A 500 µm wide, 2 µm thick Au transmission line cell and electronics providing frequencies up to 20 GHz were used for microwave measurements. The long axis of the sample was aligned perpendicularly to the microwave line, thus parallel to the excitation field h rf as indicated in the inset of Fig. 2. V ISH was measured by a lock-in technique (with the microwave power turned on and off at a frequency of a few kHz) with electrical connections through gold leads at equal distance to the area of excitation. Magnetotransport measurements of the YIG|NM slabs were performed using a 4-point configuration. The samples were placed at the center of an electromagnet, which can be rotated around its axis in order to obtain curves of magnetoresistance vs. angle. The measurement cell was placed in a cryostat, with the possibility to cool down to 77 K. All the measurements presented in this paper were performed at room temperature, except for those reported in Fig. 5.

III. EXPERIMENTAL RESULTS AND ANALYSIS

A. Inverse spin Hall voltage: YIG|Pt vs. YIG|Ta

First, we compare in Fig. 2 the inverse spin Hall voltages measured at 3.5 GHz (P = +10 dBm) in the YIG|Pt and YIG|Ta bilayers. It shows that one can electrically detect the FMR of YIG in these hybrid systems 6 . The spin current J s pumped into the adjacent normal metal by the precessing magnetization in YIG is converted into a charge current by ISHE,

J e = 2e Θ SH J s , (1)
where e is the electron charge and the reduced Planck constant. This leads to a transverse voltage V ISH (across the length of the YIG|NM slab), as sketched in the inset of Fig. 2. Moreover, V ISH must change sign upon reversing the magnetization of YIG because of the concomitant reversal of the spin pumped current J s (hence J e ). This is observed in both the YIG|Pt and YIG|Ta systems, where V ISH is odd in applied magnetic field, which shows that the voltage generated at resonance is not due to a thermoelectrical effect. The striking feature to be observed here is the opposite signs of V ISH in these two samples. This remains true at all microwave frequencies (from 2 to 8 GHz) and power levels (from -8 to +10 dBm) which were measured, as well as for the different YIG|Pt and YIG|Ta bilayers made from YIG1 and YIG2 samples. It thus confirms that the spin Hall angles in Ta and Pt have opposite signs, as predicted by ab initio calculations 24 and inferred from measurements where the spin current was generated by a metallic ferromagnet 20,23 . Moreover, from the electrical circuit which was used in the measurements (anode of the voltmeter is on the left in Fig. 2 inset), it can be found that Θ Pt SH > 0 while Θ Ta SH < 0. The precise estimation of the spin Hall angles in these two materials requires the more analysis presented in the following sections. Still, it is interesting to note that the 4 µV amplitude of V ISH measured in Fig. 2 on our 15 nm thick Pt is close to the one reported in Ref. 17 (2 to 3 µV) with comparable experimental conditions.

B. Dependence of inverse spin Hall voltage on Ta thickness

In this work, we have measured the dependence of V ISH only on Ta thickness. The study as a function of Pt thickness was already reported in Ref. 17 , using a similar 200 nm thick YIG film (fabricated in the same lab). In Fig. 3, we have plotted using red squares the dependence of V ISH on the Ta thickness measured on the series of samples described above. Here, V ISH is produced by the precession of magnetization in YIG, resonantly excited at 3.8 GHz by the microwave field (P = +10 dBm). V ISH increases from less than 2 µV up to 70 µV as the Ta layer thickness is reduced from 15 nm to 2 nm, at which the maximal voltage is measured. For the thinnest Ta layer (t Ta = 1.5 nm), V ISH drops to about 10 µV, a value close to the one observed at t Ta = 10 nm. A similar dependence of V ISH on Pt thickness was reported in Ref. 17 , where a maximum of voltage was observed between t Pt = 1.5 nm and t Pt = 6 nm.

The resistance measured across the length of the YIG|Ta slab is also plotted with green crosses in Fig. 3 as a function of t Ta (see right scale). It is interesting to note that both V ISH and R follow a similar dependence on the Ta thickness, if one excludes the thinnest Ta layer, which might be discontinuous or oxidised, and thus exhibits a very large resistance (R = 95 kΩ is out-of-range of the graph).

To analyze the thickness dependence of the inverse spin Hall voltage, we follow the approach derived in Ref. 17 . The spin diffusion equation with the appropriate source term and boundary conditions leads to the following expression:

V ISH = Θ SH G ↑↓ G ↑↓ + σ λ sd 1-exp (-2tNM/λ sd ) 1+exp (-2tNM/λ sd ) × hLP f sin 2 (θ) 2et NM (1 -exp (-t NM /λ sd )) 2 1 + exp (-2t NM /λ sd ) , (2)
where σ is the conductivity of the normal metal, t NM its thickness, L the length of the YIG|NM slab excited at frequency f by the microwave field, θ the angle of precession of YIG, and P an ellipticity correction factor. The latter depends on the excitation frequency 18 and in our case P ≃ 1.25. From Eq.2, the amplitude of V ISH depends on the transport parameters λ sd , G ↑↓ and Θ SH , as well as on the resonant precession angle θ. We do not have a direct measurement of θ, but it can be evaluated from the strength of the microwave field h rf and the measured linewidth ∆H [START_REF] Gurevich | Magnetization Oscillations and Waves[END_REF] . By performing network analyzer measurements and considering the geometry of the transmission line, we estimate the strength of the microwave field h rf ≃ 0.2 Oe for a P = +10 dBm output power from the synthesizer. For the series of YIG|Ta samples, it yields a precession angle θ ≃ 3.3 • in YIG at 3.8 GHz. Nevertheless, the measurements presented in Fig. 3 are not sufficient to extract independently G ↑↓ and Θ SH .

The thickness dependence of V ISH primarily depends on λ sd , through the argument of the exponential functions in Eq.2. The spin diffusion length can thus be adjusted to fit the shape of V ISH vs. t Ta in Fig. 3. The series of lines in Fig. 3 displays the result of calculations based on Eq.2 for three different values of λ sd , using the thickness dependent conductivity σ Ta measured experimentally. A very good overall agreement to the data is found for a spin diffusion length λ Ta sd = 1.8 nm. We explain the discrepancy observed at t Ta = 1.5 nm, at which the measured voltage is about five times smaller than predicted, by the fact that the thinnest Ta layer is discontinuous or oxidised, as already pointed out. We note that the spin diffusion length extracted from the YIG|Ta data of Fig. 3 is somewhat shorter than the 2.7 nm inferred from nonlocal spin-valve measurements 23 .

C. Magnetoresistance: YIG|Pt vs. YIG|Ta

We now turn to the measurements of dc magnetoresistance in our hybrid YIG|NM bilayers. We have measured the variation of resistance in the exact same samples as the ones studied by ISHE in Fig. 2, YIG|Pt(15 nm) and YIG|Ta(3 nm), as a function of the angle of the applied field with respect to the three main axes of the slabs. In these experiments, the applied field was fixed to H = 3 kOe (sufficient to saturate the YIG), and a dc current of a few mA together with a 6 1/2 digits voltmeter were used to probe the resistance of the NM layers in a 4-probe configuration. The results obtained by rotating the magnetic field in the plane of the sample (angle α), from in-plane perpendicular to the charge current J e to out-of-plane (angle β) and from in-plane parallel to J e to out-of-plane (angle γ) are presented in Figs. 4a, 4b, and4c, respectively (see also associated sketches).

In both the YIG|Pt and YIG|Ta bilayers, we do observe some weak magnetoresistance (∆R max /R 0 of 5 •10 -5 and 4 • 10 -5 , respectively), as it was first reported on the YIG|Pt system 22 . We checked that this weak variation does not depend on the sign or strength of the probing current. In contrast to the inverse spin Hall voltage measurements presented in Fig. 2, we also note that the sign (or symmetry) of the effect is identical in YIG|Pt and YIG|Ta.

In order to interpret this magnetoresistance, it is important to understand its dependence on all three different angles, α, β and γ, shown in Fig. 4. If one would just look at the in-plane behavior (Fig. 4a), one could conclude that the NM resistance R changes according to some anisotropic magnetoresistance (AMR) effect, as if the NM would be magnetized at the interface with YIG due to proximity effect 22 . But with AMR, R depends on the angle between the charge current J e and the magnetization (applied field H). Hence, no change of R is expected with the angle β, whereas R should vary with the angle γ, which is exactly opposite to what is observed in Figs. 4b and4c, respectively. Therefore usual AMR as the origin of the magnetoresistance in YIG|Pt and YIG|Ta bilayers has to be excluded.

Instead, the spin Hall magnetoresistance (SMR) mechanism proposed in Ref. 8 is well supported by our magnetoresistance data. In this scenario, the electrons carried by the charge current in the NM layer are deflected by SHE in opposite directions depending on their spin. Those whose spin is flipped by scattering at the interface with the FM can oppose the initial current by ISHE and lead to an increase of resistance. Therefore, the spin Hall magnetoresistance depends on the relative angle between the magnetization M of the FM and the accumulated spins s at the FM|NM interface: The increase of resistance is maximal when M and s are perpendicular, because the spin-flip-scattering governed by G ↑↓ at the interface is the largest. In the geometry depicted in Fig. 4, the charge current is applied along y, hence the spins accumulated at the YIG|NM interface due to SHE are oriented along x. The dashed lines plotted in Figs. 4a-c are the prediction of the SMR theory.

R = R 0 + ∆R max sin 2 (M , s) . (3)
As can be seen, Eq.3 explains well the presence (absence) of resistance variation upon the applied field angles α and β (γ). Due to demagnetizing effects, the magnetization of YIG is not always aligned with the applied field. This is the reason why the measured curves in Figs.4a and 4b have different shapes, and a simple calculation [START_REF] Gurevich | Magnetization Oscillations and Waves[END_REF] of the equilibrium position of M in combination with Eq.3 reproduces them quite well.

The SMR ratio was also calculated in Ref. 8 :

SMR = ∆R max R 0 = Θ 2 SH 2λ 2 sd σtNM G ↑↓ tanh 2 tNM 2λ sd 1 + 2λ sd σ G ↑↓ coth tNM λ sd . (4)
As for the inverse spin Hall voltage V ISH (Eq.2), SMR depends on all the transport parameters G ↑↓ , Θ SH and λ sd , which therefore cannot be extracted individually from a single measurement. In section IV A, we will take advantage of the combined measurements of V ISH (Figs. 2 and3) and SMR (Fig. 4) to do so. For now, it is interesting to point out that because both SHE and ISHE are at play in spin Hall magnetoresistance, the SMR depends on the square of the spin Hall angle. This explains the positive SMR for both YIG|Pt and YIG|Ta, even though the spin Hall angles of Pt and Ta are opposite. Finally, it would have been interesting to measure the dependence of SMR on Ta thickness (the dependence on Pt thickness was studied in Refs. 22 and 28 ). Unfortunately, it was difficult to realize low noise 4-point contacts to investigate the faint magnetoresistance on the series of Ta samples prepared to study V ISH vs. t Ta . From our attempts, we found that the SMR of YIG|Ta(10 nm) is less than 2 • 10 -5 . This is consistent with the decrease predicted by Eq.4 (assuming λ Ta sd = 1.8 nm) with respect to the SMR ≃ 4 • 10 -5 measured for YIG|Ta(3 nm).

IV. DISCUSSION

A. Transport parameters

As already discussed, both V ISH and SMR depend on the set of transport parameters (G ↑↓ , Θ SH , λ sd ). By studying V ISH as a function of the NM thickness, the spin diffusion length can be determined, and we found that in Ta, λ Ta sd = 1.8 ± 0.7 nm, see Fig. 3. We mention here that from a similar study on YIG|Pt, λ Pt sd = 3.0 ± 0.5 nm could be inferred 17 . This value lies in the range of spin diffusion lengths reported on Pt, which span over almost an order of magnitude 26 , from slightly more than 1 nm up to 10 nm.

There is a direct way to get the spin mixing conductance of a FM|NM interface, by determining the increase of damping in the FM layer associated to spin pumping in the adjacent NM layer 7 . Due to its interfacial nature, this effect is inversely proportional to the thickness of the FM and can be measured only on ultra-thin films. This was recently achieved in nm-thick YIG films grown by pulsed laser deposition [START_REF] Heinrich | [END_REF]31 , where spin mixing conductances G ↑↓ = (0.7 -3.5) • 10 14 Ω -1 •m -2 have been reported for the YIG|Au interface.

Even for 200 nm thick YIG films as ours, it is possible to obtain the full set of transport parameters thanks to our combined measurements of V ISH and SMR on YIG|NM hybrid structures. In fact, from Eqs.2 and 4, the ratio V 2 ISH /SMR does not depend on Θ SH , which allows to determine G ↑↓ . Then, the last unknown Θ SH TABLE I. Transport parameters obtained from the analysis of inverse spin Hall voltage (Figs. 2 and3 + Eq.2) and spin Hall magnetoresistance (Fig. 4 + Eq.4) performed on YIG|Ta(1.5 nm-15 nm) and YIG|Pt(15 nm). YIG|Ta (1.5 nm -15 nm) YIG|Pt (15 nm) σ (10 can be found from the V ISH or SMR signal. This is how we proceed to determine the transport parameters which are collected in Table I. The drawback of this method is that it critically relies on: i) λ sd , which enters in the argument of exponential functions in Eqs.2 and 4; and ii) the angle of precession θ in the inverse spin Hall experiment, since V 2 ISH /SMR ∝ θ 4 . Our estimation of θ being within ±25%, the value extracted for G ↑↓ from the ratio V 2 ISH /SMR can vary by a factor up to 8 due to this uncertainty. The spin Hall angle Θ SH is less sensitive to other parameters, still it can vary by a factor up to 3. This explains the rather large error bars in Table I. In this study, we did not determine the spin diffusion length in Pt, hence we used the range of values reported in the literature 26 .

The spin mixing conductances determined from our combined V ISH and SMR measurements on YIG|Ta and YIG|Pt bilayers lie in the same window as the ones determined from interfacial increase of damping in YIG|Au [START_REF] Heinrich | [END_REF] , from inverse spin Hall voltage in BiY 2 Fe 5 O 12 |Au and Pt 32 , and from first-principles calculations in YIG|Ag 25 . We would like to point out that despite the large uncertainty, G ↑↓ for YIG|Ta is likely less than for YIG|Pt. We note that the smaller damping measured in CoFeB|Ta compared to CoFeB|Pt was tentatively attributed to a smaller spin mixing conductance 20 .

The spin Hall angles that we report for Pt and Ta are both of a few percents. In particular, Θ Ta SH ≃ -0.02 lies in between the values determined from nonlocal spinvalve measurements (≃ -0.004) 23 and from spin-torque switching using the SHE (≃ -0.12) 20 .

The main conclusion which arises from the summary presented in Table I is that the sets of transport parameters determined for the hybrid YIG|Ta and YIG|Pt systems are quite similar. Apart from the opposite sign of Θ SH in Ta and Pt, the main difference concerns the conductivity: σ β-Ta is roughly one order of magnitude smaller than σ Pt . This explains the large inverse spin Hall voltages that can be detected in our YIG|Ta bilayers (up to 70 µV at P = +10 dBm), since from Eq.2 V ISH ∝ 1/σ, which could be a useful feature of the Ta layer.

B. Influence of a dc current on FMR linewidth

Onsager reciprocal relations imply that if there is an ISHE voltage produced by the precession of YIG, there must also be a transfer of spin angular momentum from the NM conduction electrons to the magnetization of YIG, through the finite spin mixing conductance at the YIG|NM interface 25 . Therefore, one would expect to be able to control the relaxation of the insulating YIG by injecting a dc current in an adjacent strong spin-orbit metal, as it was shown on YIG|Pt in the pioneering work of Kajiwara et al. 6 . Although this direct effect is well established when the ferromagnetic layer is ultra-thin and metallic [33][34][35][36] , only a few works report on conclusive effects on micron-thick YIG 6,37,38 or provide a theoretical interpretation to the phenomenon 39 .

The 200 nm thick YIG films that have been grown for this study are about 6 times thinner than the one used in Ref. 6 , with an intrinsic relaxation close to bulk YIG. Because the spin transfer torque is an interfacial effect and sizable spin mixing conductances have been measured in our YIG|Ta and YIG|Pt bilayers, our samples must be good candidates to observe the direct effect of a dc current on the relaxation of YIG. Due to their large resistance, β-Ta films are not convenient to pass the large current densities required to observe such an effect (large Joule heating). Therefore, we have conducted these experiments only on the YIG|Pt films prepared in this work.

The inverse spin Hall voltage measurements presented in Fig. 2 have therefore been repeated in presence of a dc current flowing through the Pt layer. This type of experiment, where a ferromagnetic layer is excited by a small amplitude signal and a spin polarized current can influence the linewidth of the resonance, has already been reported on spin-valve spin-torque oscillators 40,41 and NiFe|Pt bilayers 42 . The results obtained on our YIG(200 nm)|Pt(15 nm) at 77 K when the dc current is varied from -40 to +40 mA are displayed in Fig. 5.

Let us now comment on these experiments. We first emphasize that the current injected in Pt is truly dc (not pulsed). A sizable Joule heating is thus induced, as reflected by the increase of Pt resistance. As a consequence, the main effect of dc current injection at room temperature is the displacement of the resonance towards larger field, due to the decrease of the YIG saturation magnetization M s . To avoid this trivial effect, we have performed these experiments directly in liquid nitrogen. In that case, the increase of Pt resistance is very limited (+0.2% at ±40 mA). We note that when cooled from 300 K down to 77 K, the peak of the inverse spin Hall voltage measured in the YIG|Pt bilayer is displaced towards lower field due to the increase of M s of YIG (from 140 emu/cm 3 up to 200 emu/cm 3 ), and its amplitude slightly decreases.

The main conclusion that can be drawn from Fig. 5 is that there is basically no effect of the dc current injected in Pt on the YIG resonance. We stress that the maximal current density reached in Pt in these experiments is J e = 2.4 • 10 9 A.m -2 , i.e., twice larger than the one at which YIG magnetization oscillations were reported in Ref. 6 . In our experiments, we are not looking for autooscillations of YIG, which requires that the damping is fully compensated by spin transfer torque, but only for some variation of the linewidth. The fact that we do not see any change in the shape of the resonant peak of our 200 nm thin YIG film is thus in contradiction with the observation of bulk auto-oscillations in thicker films 6 .

We have also performed similar experiments on the other YIG(200 nm)|Pt samples which were prepared using the two different YIG films grown for this study. Although the current density was increased up to 6 • 10 9 A.m -2 , we were never able to detect any sizable variation of the linewidth of YIG. Instead, we have measured that the dc current can affect the inverse spin Hall voltage in different ways. First, when a charge current is injected into Pt, a non-zero offset of the lock-in signal can be detected (it was subtracted in Fig. 5). This is due to the increase of Pt resistance induced by the microwave power, as it was verified by monitoring this offset while varying the modulation frequency of the microwave. Secondly, the amplitude of the V ISH peaks can be affected by the dc current (but again, not the linewidth). This effect can at first be confused with some influence on the relaxation of YIG, because it displays the appropriate symmetries vs. field and current. But instead, we have found that this is a bolometric effect 43 : when the YIG is excited at resonance, it heats up, thereby heating the adjacent Pt whose resistance gets slightly larger. Hence an additional voltage to V ISH is picked up on the lock-in due to the non-zero dc current flowing in Pt. Therefore, one should be very careful in interpreting changes in inverse spin Hall voltage as the indication of damping variation in YIG. Finally, we observed that at very large current density, the resonance peak slightly shifts towards larger field due to Joule heating, even at 77 K.

V. CONCLUSION

In this paper, we have presented and analyzed a comparative set of data of inverse spin Hall voltage V ISH and magnetoresistance obtained on YIG|Pt and YIG|Ta bilayers. We have detected the voltages generated by spin pumping at the YIG|Pt interface (already well established 6 ) and at the YIG|Ta interface (for the first time). Their opposite signs are assigned to the opposite spin Hall angles in Pt and Ta 24 . From the thickness dependence of V ISH , we have been able to obtain the spin diffusion length in Ta, λ Ta sd = 1.8 ± 0.7 nm, in reasonable agreement with the value extracted from non-local spin valve measurements 23 . From symmetry arguments, we have shown that the weak magnetoresistance measured on our hybrid YIG|NM layers cannot be attributed to usual AMR, but is instead well understood in the framework of the recently introduced spin Hall magnetoresistance (SMR) 8 . By taking advantage of the combined measurements of V ISH and SMR performed on the same samples, we have been able to extract the spin Hall angles in Pt and Ta, as well as the spin mixing conductances at the YIG|Pt and YIG|Ta interfaces.

These transport parameters have all been found to be of the same order of magnitude as those already measured 20,[START_REF] Heinrich | [END_REF] or predicted 25 . We believe that at least part of the discrepancies between the parameters evaluated in different works 26 depend on the details of the YIG|NM interface 31 and on the quality of the NM 18,19,23 .

Finally, we could not detect any change of linewidth in our YIG|Pt samples by passing large current densities through the Pt layer. One might argue that our high quality 200 nm YIG thin films are still too thick to observe any appreciable effect of spin transfer torque, which is an interfacial mechanism, or that the spin-waves which can auto-oscillate under the action of spin transfer at the interface with Pt are different from the uniform mode that we excite with the microwave field in our experiments 6,39 . If one would estimate the threshold current required to fully compensate the damping of all the magnetic moments contained in our YIG films 20,39 , J th ≃ 2eαωM s t YIG /(Θ SH γ ), one would get current densities of about 10 11 A.m -2 . This is 20 times larger than the largest current density which we have tried. Thus the lack of a visible effect in our Fig. 5 is not a real surprise in itself, but it is inconsistent with the results reported in Ref. 6 . Future experiments on ultra-thin YIG|NM hybrid films, in which the spin mixing conductance can be directly determined from the interfacial increase of damping [START_REF] Heinrich | [END_REF] , might give a definite answer to this point.

FIG. 1 .

 1 FIG. 1. (Color online). (a) Standard in-plane FMR spectrum of a bare YIG 200 nm thin film used in this study. (b) Full FMR linewidth vs. frequency.

FIG. 2 .

 2 FIG. 2. (Color online). Inverse spin Hall voltage measured at 3.5 GHz for YIG|Ta and YIG|Pt. Inset: sketch of the experiment.

FIG. 3 .

 3 FIG. 3. (Color online). Dependence of inverse spin Hall voltage on Ta thickness (red squares, left scale). The microwave frequency is 3.8 GHz (P = +10 dBm). The lines are theoretical predictions 17 from Eq.2 for different values of λ sd , with the parameters G ↑↓ = 4.3 • 10 13 Ω -1 •m -2 and ΘSH = -0.02. The resistance of the samples is also displayed (green crosses, right scale). The resistance of the 1.5 nm thin Ta sample (95 kΩ) is out-of-range.

FIG. 4 .

 4 FIG. 4. (Color online). (a-c) Magnetoresistance in YIG|Ta and YIG|Pt as a function of the angle of the applied field (H = 3 kOe) sketched at the top (the samples are the same as the ones measured in Fig.2). Dashed lines are predictions from Eq.3 of the SMR theory 8 .

FIG. 5 .

 5 FIG. 5. (Color online). Inverse spin Hall voltage measured at 2.95 GHz (P = +10 dBm) for YIG|Pt as a function of the dc current flowing in the Pt layer. A small current dependent offset (< 0.2 µV) has been subtracted to the data.

  6 Ω -1 •m -1 ) 0.08 -0.75 2.45 ± 0.10 λ sd (10 -9 m) 1.8 ± 0.7 n/a [from 1.5 to 10] 26 G ↑↓ (10 13 Ω -1 •m -2 )

		4.3± 11 2	6.2± 14 4
	ΘSH	-0.02± 0.008 0.015	0.03± 0.04 0.015

ACKNOWLEDGMENTS

This research was supported by the French ANR Grant Trinidad (ASTRID 2012 program).


Flavio Caresana

Gabriele Comodi
email: g.comodi@univpm.it

Leonardo Pelagalli

Massimiliano Renzi

Sandro Vagni

Use of a test-bed to study the performance of micro gas turbines for cogeneration applications

Keywords: micro gas turbine, cogeneration, test-bed, thermal recovery, electrical performance, primary energy saving

This paper presents the experimental results of a test-bed developed to analyze a 100 kW cogeneration micro gas turbine (MGT). Its main feature, flexibility, allows MGT testing in nearly all the conditions that final use may entail, in particular simulation of operating conditions and thermal loads in broadly different electrical power, flow rate and temperature conditions of the cogeneration circuit. The MGT and its operating modes are described and the test-bed design philosophy is outlined. Finally the methods applied to test the cogeneration plant's performance are described and the experimental results in terms of power, efficiency, primary energy saving (PES) index and emissions are reported.

In the tests electrical power ranged between 40 and 100 kW; electrical efficiency reached ~ 29 % in the 80-100 kW range. Analysis of the performance of the recovery heat exchanger at different values of inlet and outlet water temperature and flow rate showed that the maximum thermal power recovered was ~ 160 kW and the maximum PES index was about 24 %. With regard to the main pollutants at the exhaust, NO x concentrations were < 6 ppmv in the whole MGT working range, while CO concentrations increased substantially at low electrical power values, starting from around 10 ppmv at nominal load.

Introduction

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In conventional energy systems electrical power is conveyed from large-scale plants located far away from the region where it is used, while energy for heating is supplied separately as fuel on a local basis. In the future distributed energy networks are likely to provide both electrical and heating energy through CHP plants built close to the consuming region [START_REF] Pepermans | Distributed generation: definition, benefits and issues[END_REF][START_REF]IEA distributed generation in liberalised electricity markets[END_REF]. The value of MGTs as components of distributed energy systems lies in their low environmental impact in terms of pollutants. MGTs appear to be particularly well suited for service sector, household and small industrial applications [START_REF] Soares | Microturbines[END_REF][START_REF] Pilavachi | Mini-and micro-gas turbines for combined heat and power[END_REF][START_REF] Macchi | Natural gas micro-cogeneration[END_REF][START_REF] Zogg | Using microturbines for distributed generation[END_REF].

Despite this, the market for MGTs is still modest; research is therefore being conducted to make them more appealing through application of concepts such as trigeneration [START_REF] Campanari | Microturbines and trigeneration: optimization strategies and multiple engine configuration effects[END_REF][START_REF] Bruno | Performance analysis of combined microgas turbines and gas fired water/LiBr absorption chillers with post-combustion[END_REF][START_REF] Hwang | Potential energy benefits of integrated refrigeration system with microturbine and absorption chiller[END_REF], inlet air cooling [START_REF] Hwang | Potential energy benefits of integrated refrigeration system with microturbine and absorption chiller[END_REF][START_REF] Brandon | Inlet air supercharging and low NOx duct burners[END_REF][START_REF] Zaltash | Effect of microturbine inlet air cooling and humidification on IES performance[END_REF], biogas use [START_REF] Lymberopoulos | Micro turbines and their application in bio-energy, Centre for renewable energy sources[END_REF][START_REF] Blasco | Microturbinas: utilización con biogas de digestión y vertederos, Bioprom Project Workshop[END_REF] or combinations thereof [START_REF] Bruno | Integration of absorption cooling systems into micro gas turbine trigeneration systems using biogas: Case study of a sewage treatment plant[END_REF]. Some authors have also studied the possibility of enhancing performances by adding a bottoming organic Rankine cycle to recover the thermal power of exhaust gases [START_REF] Invernizzi | Bottoming micro-Rankine cycles for micro-gas turbines[END_REF][START_REF] Caresana | Micro combined plant with gas turbine and organic cycle[END_REF].

The majority of the relevant works on MGT cogeneration units report operation and emissions data at nominal load in ISO conditions [17]. Most of these data are taken directly from the manufacturer's documentation, whereas few studies have assessed non-nominal and/or partial load conditions. For example, Colombo et al. [START_REF] Colombo | Experimentation on a cogenerative system based on a microturbine[END_REF] describe a cogeneration test-bed and the performance and pollutant emissions of a 100 kW MGT at partial load using a set outlet temperature of the heat recovery boiler water. Fairchild et al. [START_REF] Fairchild | Experimental and theoretical study of microturbinebased BCHP system[END_REF] report the experimental results of a 30 kW cogeneration MGT, also running at partial load, focusing on the influence of backpressure on performance; however, their thermal performance data are not for the turbine, but for the whole HVAC system at the site where it is installed.

The present paper describes MGT performance in partial electrical and thermal load conditions, providing experimental data from an ad hoc test-bed. density of the water in the primary circuit at the RHE inlet (kg/m 3 )

Nomenclature

The cogeneration micro gas turbine

The machine chosen for the tests was a Turbec T100 PH cogeneration MGT based on a regenerative Brayton cycle; it is a more recent version of the one investigated by Colombo et al. [START_REF] Colombo | Experimentation on a cogenerative system based on a microturbine[END_REF] and features an embedded fuel compressor. The unit is fired with natural gas and the combustor, a lean-premixed-low-emission type, guarantees volumetric concentrations of NOx and CO < 15 ppmv at 15% O 2 at full load.

In ISO 2314 [17] conditions the machine produces 100 kW of electrical power with an efficiency of 30%; the combustion products exit the turbine at a temperature of 645 °C, cross the regenerator and enter the recovery heat exchanger at 270 °C. The RHE is a counter-flow type designed for a water mass flow rate of 2 l/s and temperatures of 50 °C and 70 °C at the inlet (to avoid dew condensation in the worst foreseeable conditions) and outlet, respectively. At these nominal conditions stated thermal power production is ~ 155 kW, with a thermal efficiency of 47 %.

An electrically driven bypass valve fitted along the RHE inlet manifold can divert some of the exhaust gases to the chimney, to regulate the thermal power recovered.

The MGT performance data as supplied by the manufacturer are summarized in Table 1.

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In practice, demand is represented by user-set values of electrical power and fluid temperature at the RHE outlet, and by selection of the operating mode (electrical or thermal priority).

In the electrical priority mode the MGT produces the electrical power required by the user and the fluid temperature is kept at the set-point through regulation of bypass valve opening.

In the thermal priority mode the bypass valve is completely closed and all the exhausts pass through the RHE to produce thermal power. The temperature of the hot fluid is kept at its set-point by variations in the amount of electricity produced; the user-selected electrical power value is the threshold.

Electrical priority mode operation with the bypass valve completely closed is equivalent to operation in the thermal priority mode.

Test-bed

The test-bed, illustrated in figure 1, was designed to simulate a broad range of thermal demands and is therefore capable of testing the cogeneration system in a wide range of user-selected conditions, involving several combinations of electrical power, water flow rate and operating temperatures.

The test-bed consists of four main subsystems: the natural gas feeding plant, the inlet and extraction manifold, the grid connection interface, and the hydraulic system for thermal power dissipation.

The main components of the natural gas feeding plant are a valve that reduces the grid pressure to the required turbine inlet value and a fuel flow meter measuring the natural gas mass flow rate.

The inlet and extraction manifold consists of three ducts and a centrifugal extractor fan. One duct provides inlet air that serves both as a working fluid in the turbo compressor and as a coolant for the auxiliary system; the other two ducts collect cooling air from the turbine enclosure and the exhausts at the RHE outlet. All ducts are fitted with thermocouples; a gas analyzer for emission control is installed in the exhaust gas duct.

The grid connection interface manages the parallel connection with the grid and measures the electrical power produced. Its main components are a safety switch, to disconnect the turbine from the grid in case of emergency, and a grid analyzer to measure electrical parameters and power production.

The hydraulic system for thermal power dissipation is the most complex part of the test-bed. It consists of two separate circuits thermally interconnected by a plate heat exchanger. The primary circuit contains the water circulating through the turbine RHE, while the secondary circuit contains a water-glycol mixture that discharges the heat to the outside air by circulating through an air cooler.

M A N U S C R I P T A C C E P T E D ACCEPTED MANUSCRIPT 5
The water circuit is provided with a volumetric flow meter and with two Pt100 temperature sensors at the RHE inlet and outlet.

Valves V1 and V2 in the water circuit allow flow rate regulation through the RHE. V1 is only manually operated, whereas V2-which keeps the water temperature at the RHE outlet at its set-point-can also be controlled remotely by a PID.

The pneumatic mixing valve V3, installed in the secondary circuit to keep the water temperature at the RHE inlet at its set-point, is automatically driven by a PID controller.

All measurement instruments and control signals are managed by means of a PLC connected to a PC.

The measurement devices installed and their accuracy are listed in Table 2.

Evaluation of micro gas turbine performance: results and comments

The test-bed's characteristics make it possible to assess the electrical and thermal performance of the cogeneration unit in various working conditions, including electrical power settings ranging from minimum to nominal value and thermal power settings ranging from the thermal priority mode maximum to the electrical priority mode minimum. Given the dependence of MGT performance on ambient temperature [20], only experimental data acquired at nearly constant ambient temperature (from 5 to 10 °C) were considered.

First of all, the tests were carried out with the water volumetric flow rate and the inlet temperature in the primary circuit at nominal values (2 l/s and 50 °C, respectively), using several different combinations of electrical power and degree of bypass valve opening (from fully open to fully closed).

These data, reported in section 4.1, were used to draw up a performance chart of the cogeneration unit in the 40-100 kW electrical power range.

Since the MGT can be fed water at non-nominal conditions, depending on user requirements, additional tests were carried out with different water flow rates and/or operating water temperatures. These data are reported in sections 4.2 and 4.3.

For each working point the thermal power recovered and the electrical, thermal and total efficiency of the unit were calculated as:

( ) This suggests the need for installing a heat dissipator in real applications, to prevent water boiling in the RHE in the event of the thermal demand falling under the fully-open lower limit (PTH% <15 %). When the turbine is used only as an electrical generator for an extended period of time, the RHE should be drained.

w_in w_out w w w th T T c Q ρ P - = 1) LHV m P η f e e & = 2) M A N U S C R I P T A C C E P T E D ACCEPTED MANUSCRIPT 6 LHV m P η f th th & = 3) th e f th e tot η η LHV m P P η + = + = & 4)

Cogenerator performance at nominal inlet thermal conditions

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The total efficiency data are reported in the performance chart next to each experimental point; the corresponding outlet temperature, T w_out , can be read on the secondary axis on the right.

The total efficiency data are also reported in figure 3 with the electrical efficiency data. For each working point, thermal efficiency is the difference between total and electrical efficiency.

The average electrical efficiency is 28.5 % at maximum load, peaking at about 29.5 % for a net electrical power of 80 kW. Furthermore, efficiency declines to an average of 23.5 % at 40 kW. Taking into account inlet and exhaust duct pressure losses, these values are consistent with those stated in the manufacturer's data sheet [17].

Figure 3 shows that the electrical efficiency does not depend on the position of the bypass valve (i.e. on operating mode); rather, for each electrical power setting the lower the thermal power demand (i.e. the greater the bypass valve opening), the lower the thermal efficiency (and thus total efficiency). For example, at 100 kW of electrical power, total efficiency is 74 % with the valve fully closed (thermal priority) and 36 % with the valve fully open (see also the data in figure 2).

At the same time, for a given level of thermal power demand the greater the electrical power produced, the lower total efficiency. For example, it can be seen from figure 2 that the unit produces about 80 kW of thermal power either working in thermal priority mode, at 40 kW of electrical power, or working in electrical priority mode, at 100 kW of electrical power with the bypass valve about 1/3 open. In the former case total efficiency is around 74 %, in the latter it is slightly more than 51 %.

Finally the concentrations of the primary pollutants [START_REF] Soares | Gas Turbines: A Handbook of Air, Land and Sea Applications[END_REF] at 15 % O 2 dilution are reported in the performance chart. NO x were low in the whole electrical working range of the MGT, with a peak of ~6 ppmv in correspondence with the machine's minimum load; CO concentrations < 10 ppmv were measured at 100 and 80 kW, while at lower electrical power settings concentrations were significantly higher, with a maximum of 1550 ppmv at 40 kW.

Cogenerator performance at nominal electrical power

In this section the data show how water parameters at the RHE influence the recoverable thermal power when the system is working with the bypass valve fully closed (i.e. in thermal priority mode).

For each electrical power setting, the recoverable thermal power depends on the water parameters, because whereas modification of inlet water parameters does not change the temperature and amount of the exhausts Here we focus on the 100 kW nominal electrical power point; the next section addresses lower power values.

Two experimental data sets are compared in figure 4, one regarding the RHE fed with the nominal mass flow rate of water at different temperatures and the other regarding the RHE fed with different mass flow rates at the nominal inlet temperature. The parameters (thermal power recovered, mass flow rate and inlet temperature) are reported as a function of outlet temperature.

As shown by the tendency lines, the data sets of the two cases (represented respectively as circles and squares in figure 4) converge at the nominal thermal values of Q w = 2 l/s, T w_in = 50 °C, and T w_out = 70 °C.

From this point, if the water temperature at the outlet is increased from the nominal value, the inlet temperature is kept constant at 50 °C, and the mass flow rate is reduced accordingly (squares), more thermal power is recovered with respect to the other case (circles).

Alternatively, if the water outlet temperature is increased from the nominal value, keeping the mass flow rate constant at 2 l/s, the inlet temperature rises accordingly (circles); as a result the mean temperature difference between the MGT exhausts and the water is lower than in the previous case and less thermal power is recovered.

Cogenerator performance at nominal water inlet and outlet temperatures

In this section the data show how water conditions influence the thermal priority tendency lines. The data of the two cases converge at the nominal values. Starting from the nominal point, if the electrical power is reduced while keeping the inlet and outlet temperature at 50 °C and 70 °C, respectively (squares in figure 5), less thermal power is recovered with respect to the other case (circles). Alternatively, if the electrical power is reduced while keeping the mass flow rate constant at 2 l/s and the inlet temperature at 50 °C (circles), the temperature at the RHE outlet decreases; the mean temperature difference between the MGT exhausts and water is greater than in the previous case, and more thermal power is recovered accordingly.

M A N U S C R I P T A C C E P T E D

Primary energy saving

To complete the analysis of the cogeneration unit this section reports some data relating to the comparison with a conventional separate generation system, which are useful to evaluate the practical use of the cogeneration unit.

From the energy point of view, this is done through the PES index, which quantifies the amount of primary energy that can be saved by heat and power cogeneration compared with separate generation, as:

th_ref th e_ref e η η η η 1 1 PES + - = 6)
where η e_ref and η th_ref are the reference electrical and thermal efficiencies of the separate generation systems.

If these are assumed to be η e_ref = 40 % and η th_ref = 80 % respectively, as prescribed by the Italian Authority for Electrical Energy and Gas, the MGT cogenerator guarantees values of the PES index between 0.19 and 0.23 when working in thermal priority mode (bypass valve fully closed; see figure 6). Nearly all the working points in the electrical priority mode fall in the negative PES area, since the greater the bypass valve opening the lower the PES, i.e. more primary energy is consumed by the cogenerator than by the separate generation systems.

Conclusions

This paper reports the experimental results of a 100 kW cogeneration MGT obtained with a test-bed designed to simulate a broad range of operating conditions including partial electrical and thermal loads.

The main performance parameters of the cogenerator are first reported on a single chart drawn up for the nominal thermal conditions; the emissions data are reported as tables.

Some working points at non-nominal conditions are then analyzed, both to highlight the test-bed capabilities and to provide insights into and information on the MGT working modes. Analysis of different electrical power settings between 40 and 100 kW yielded a maximum electrical efficiency of ~ 29.5% at 80 kW and a minimum of ~ 23.5% at 40 kW.

As regards pollutants, NO X concentrations were very low in all working conditions, while CO concentrations were low at high-medium power and rose steeply at lower loads.

The tests showed heat priority operation to be the most efficient mode, as reflected both in total efficiency and in PES index values, and cogeneration with electrical priority to be the least efficient mode, due to bypass valve M A N U S C R I P T A C C E P T E D ACCEPTED MANUSCRIPT 10 opening, hence to the amount of heat discharged. A negative PES was actually found in a number of conditions with the bypass valve not closed, demonstrating that the electrical priority mode may be less efficient than separate production. The maximum thermal power recovered was about 160 kW, with a thermal and a total efficiency around 44 % and 74 %, respectively.

When the cogenerator was compared with standard separate production with η e_ref = 40 % and η th_ref = 80 % the maximum PES index was about 0.23.

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RESEARCH HIGHLIGHTS:

• The test-bed allowed analyzing the MGT in nearly all cogenerative conditions. • Electrical efficiency reached ~ 29 % in the 80-100 kW range.

• The maximum thermal power recovered was ~160 kW (thermal efficiency ~44 %).

• Heat priority guaranteed the highest total efficiency (~74 %) and PES (~ 0.23).

• NO x were < 6 ppmv, CO were ~10 ppmv at nominal load and increased at low power.

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Figure 2

 2 Figure 2 reports the cogenerator's electrical and thermal performances obtained at different power settings between 40 and 100 kW (using 20 kW steps) with constant water volumetric flow rate and inlet temperature in the primary circuit (2 l/s and 50 °C, respectively).At each given electrical power setting, the thermal power produced was highest with the machine working in thermal priority mode. When the machine was switched to the electrical priority mode, opening of the bypass valve reduced the water temperature at the RHE outlet, hence the thermal power recovered. Tests were carried out with the bypass valve in 1/3, 2/3 and fully open positions.The dotted lines in figure2join the points sharing the same percent thermal power recovered (PTH%): ( ) ( ) e 0 th e th P P P P 100 PTH% ⋅ =

  , it does alter the outlet temperature of both water and exhausts, hence the mean temperature difference between the two fluids.

Figure 5

 5 Figure 5 reports the data obtained with water fed at the nominal inlet temperature (50 °C) with a constant temperature of 70 °C set at the RHE outlet. Starting from P e = 100 kW and nominal thermal conditions, the electrical power was reduced to 70 kW (by 5 kW steps) and the water mass flow passing through the RHE was reduced accordingly. The data used in figure 2 are also reported, to allow direct comparison.

Figure 1 .Figure 2 .Figure 3 .Figure 4 .Figure 6 .

 12346 Figure 1. Layout of the MGT cogeneration test-bed Figure 2. MGT performance chart in nominal inlet thermal conditions Figure 3. Electrical and total efficiency in nominal inlet thermal conditions Figure 4. MGT thermal performance at nominal electrical power Figure 5. MGT performance at nominal water inlet and outlet temperatures Figure 6. PES index value of the MGT cogeneration plant; η e_ref = 40 % and η t_ref = 80 %

  w = 2 l/s, T w_in = 50°C T w_in = 50°C, T w_out =

Table captions Table 1 .

 captions1 Micro gas turbine performance in nominal conditions a

Table 2 .

 2 Measurement instruments

	Figure captions

Acknowledgements

This work was supported by the Italian Environment Ministry and by the Regional Government of Marche (Ancona, Italy) within the framework of project "Ricerche energetico-ambientali per l'AERCA di Ancona, Falconara e bassa valle dell'Esino".

We are grateful to Dr. Silvia Modena for the language review.


F Caresana
email: f.caresana@univpm.it

M Bilancia

C M Bartolini

Numerical method for assessing the potential of smart engine thermal management: application to a medium-upper segment passenger car

Keywords: Engine thermal management, cooling system, fuel consumption reduction, heat transfer coefficient, nucleate boiling NOMENCLATURE

Significant reductions in vehicle fuel consumption can be obtained through a greater control of the thermal status of the engine, especially under partial load conditions. Different systems have been proposed to implement this concept, referred to as improved engine thermal management. The amount of fuel saved depends on the components and layout of the engine cooling plant and on the performance of its control system. In this work, a method was developed to calculate the theoretical minimum fuel consumption of a passenger car and used as a reference in comparing different engine cooling system concepts. A high-medium class car was taken as an example and simulated on standard cycles. Models for power train and cooling system components were developed and linked to simulate the vehicle. A preliminary analysis of the engine was performed using AVL"s Boost program. The fuel consumption of the complete vehicle, equipped with a conventional cooling plant, was determined on standard cycles and compared with that of a vehicle equipped with a "perfect" cooling system, to calculate the theoretical reduction in fuel consumption.

The heat dissipation capacity of conventional engine cooling systems is designed for maximum power operating conditions. However, since vehicle engines most often operate under partial load conditions, such systems are in fact

M A N U S C R I P T A C C E P T E D ACCEPTED MANUSCRIPT
oversized, entailing unnecessary heat losses, mechanical losses, as well as parasitic losses from high rotational speed operation of the mechanical components. Adapting the cooling system to variable operating conditions can effectively reduce fuel consumption (thus curbing harmful emissions) as well as improve engine efficiency. Several advanced automotive cooling systems have been devised to achieve this goal. Most entail replacing the standard wax thermostat valve, mechanical water pump and radiator fan with a variable position smart valve, a variable-speed electrical pump and a variable-speed radiator fan.

Engine over-and under-cooling are easily managed by controlling the coolant mass flow through engine and radiator via smart valve position and/or pump speed. The enabling technology to exploit their potential is offered by computer control.

Dynamic models for the thermal system, smart thermostat valve, coolant pump, and radiator fan have been described in [START_REF] Setlur | An Advanced Engine Thermal Management System: Nonlinear Control and Test[END_REF], where the cooling system configuration has been tested by replacing the engine with an electronic immersion heater and by operating the actuators via a non-linear tracking controller. Enhancement of gasoline and diesel engine thermal management systems through a smart thermostat valve and a variable speed water pump has been discussed in [START_REF] Setlur | Nonlinear Controller for Automotive Thermal Management Systems[END_REF],

where a non-linear tracking controller has been proposed to ensure that engine temperature follows the desired trajectory. A non-linear controller has also been designed for transient temperature tracking in an experimental system involving a variable position smart valve, a variable speed electrical water pump, a variable speed electrical radiator fan, an engine block, and various sensors [START_REF] Salah | Nonlinear-Control Strategy for Advanced Vehicle Thermal-Management Systems[END_REF]; in this system a steam-based heat exchanger mimics the heat generated by the engine"s combustion process. The power consumption associated with four different simulated solutions has been compared in a comprehensive assessment of different configurations for a heavy-duty diesel engine [START_REF] Staunton | Assessment of Advanced Thermal Management Systems for Micro-Hybrid Trucks and Heavy Duty Diesel Vehicles[END_REF].

Since pure electrical solutions involve greater electricity consumption, which raise serious problems with standard lowvoltage power circuits [START_REF] Staunton | Assessment of Advanced Thermal Management Systems for Micro-Hybrid Trucks and Heavy Duty Diesel Vehicles[END_REF][START_REF] Torregrosa | Assessment of the influence of different cooling system configurations on engine warm-up, emissions and fuel consumptions[END_REF], hydraulic-driven actuators have also been proposed [e.g. 6].

Although the main goals of an improved thermal management system are greater engine efficiency and a reduction of harmful emissions, the typically faster engine heating and higher engine temperature of automotive applications result in greater comfort and safety in winter (better cabin climate control and faster windshield defrosting) [START_REF] Guzzella | Introduction to Modeling and Control of Internal Combustion Engine Systems[END_REF].

The advantages of using a smart thermal management system have also been investigated by modelling the system by means of 1-D simulation packages. Of the several commercial software packages designed for modelling an automotive cooling system most use a 1-D lumped parameter approach. Tumus et al [START_REF] Tumus | e-Thermal: A Vehicle-Level HVAC/PTC Simulation tool[END_REF] employed this approach to develop a thermal management system modelling tool. In some studies the components have not been physically built or tested, and mathematical models have been constructed to model theoretical systems (e.g. [START_REF] Luptowski | Development of the Enhanced Vehicle and Engine Cooling System Simulation and Application to Active Cooling Control[END_REF], [START_REF] Yoo | An Engine Coolant Temperature Model and Application for Cooling System Diagnosis[END_REF] and [START_REF] Eberth | Modeling and Validation of Automotive "Smart" Thermal Management System Architectures[END_REF]).

Shortening the engine starting and warm-up time can be crucial to meeting new requirements and regulations, since these phases involve the greatest fuel consumption and pollutant emissions. These questions were addressed in [START_REF] Kreith | Principles of Heat Transfer[END_REF], which modelled the basic engine cooling system and then simulated and compared various design options.

The works cited above demonstrate experimentally or simulate the benefits of flexible control cooling systems only in particular configurations. We opted against addressing a particular arrangement; instead, we decided to focus on a question that published data leave unanswered: the maximum fuel economy that can be achieved by fitting a smart thermal management system in a car. The answer should be obtained from comparisons between conventional and smart cooling systems in standard tests and would depend on the class of the car, the test cycle chosen as the reference, and on "how smart" the cooling system was. Result generalization would require analyzing data for different classes of cars and different cooling system concepts; however, the field is too new and the task too daunting to be accomplished by a single research group. Besides, obtaining such results experimentally would be quite difficult and expensive. Our contribution to answering the question is therefore through the development of a method based on numerical simulation, which we then tested in an upper-medium segment passenger car.

The method is based on defining a "perfect" cooling system, whose performance is then used as a benchmark to which different concepts can be compared. AVL"s Boost program, a well-tested commercial tool, was used to simulate the engine"s thermal and mechanical behaviour, then a code was developed in Matlab-Simulink incorporating the Boost results and the models for the cooling system, the drive train and the vehicle aerodynamics.

The main concepts behind the paper are outlined in section 2, where a "perfect" cooling system is defined. The theoretical aspects and the methods used to solve individual problems are then described in sections 3 and 4; in particular, section 3.1 deals with engine simulation, section 3.2 with the cooling system models, and section 3.3 with the vehicle characteristics and driving cycle simulator. Section 4 reports some examples of the results that can be obtained using the proposed method; finally, the conclusions are drawn in section 5.

THE "PERFECT" COOLING SYSTEM

The fuel economy that can be achieved in a vehicle powered by an internal combustion engine using a smart thermal management system depends on the cooling system"s layout, component type and performance, and control technique.

Since different arrangements may involve quite different fuel consumption performances in relation to different complexity and cost, it would be useful to know how distant a particular solution is from the optimum. This can be done by defining the performance of a perfect cooling system to which different concepts of smart cooling can then be compared.

Accordingly, the first aim of this paper was to define the characteristics of such perfect cooling system and use it as a benchmark to evaluate the fuel economy potential of a smart thermal management.

M A N U S C R I P T A C C E P T E D ACCEPTED MANUSCRIPT

A perfect cooling system is one devoid of defects, something that cannot be obtained in practice; this system was therefore used as a reference.

Definition of the perfect cooling system

A perfect cooling system is one that meets the following three principles (principle #3a may be an alternative to #3):

1) it is able to maintain the engine constantly at its set point temperature 2) it does not require additional energy for component functioning with respect to a conventional system 3) it is able to warm-up the engine instantly 3a) during the warm-up phase the system is adiabatic.

Principle #1 entails a perfect control system. Principle #2 allows neglecting the performance details of the cooling system components at this stage of the work; in fact, a reduction of the energy required for engine cooling can be assumed to be a consequence, for example, of the use of more efficient pumps or pumping strategies, but since many concepts also involve substitution of mechanical with electrical components, the gain is not guaranteed and is system-dependent.

Principles #3 and #3a make reference to the engine"s warm-up which, together with efficiency and fuel combustion, is very important for pollutant emissions. Shortening the warm-up time has been proved to reduce emissions.

Besides the mass and the thermal characteristics of the engine itself and the engine CPU strategy, the warm-up time depends on the amount of cooling fluid contained in the plant, on its flow rate, and on the system"s thermal insulation.

In fact, part of the heat generated during warm-up is transferred to the cooling fluid and part is dissipated to the ambient through the engine walls and the cooling system components.

Since principle #3 is impracticable, the more realistic principle #3a was tested as an alternative. The data obtained by applying principle #3 quantify the system"s performance without taking into consideration the effect of the warm-up transient. Differences in performance related to principles #3 and #3a allow highlighting the influence of the warm-up phase.

ENGINE AND COOLING SYSTEM THERMAL MODELLING

Engine model

To analyze how the cooling system design affects engine performance, a four-stroke, four-cylinder, naturally aspirated gasoline engine was simulated using AVL"s Boost program. The engine"s main technical data are reported in table 1.

The Boost diagram and the simulated performances of the engine in Wide Open Throttle (WOT) condition are shown in figure 1.

Complete engine behaviour simulations were run at different thermal levels, ranging from cold to fully warmed up. In the Boost program the different thermal levels were simulated by taking different values for the liner mean temperature w T , which for the present engine ranges from 107°C (fully warmed up) to 25°C (cold).

For each thermal level the simulations consisted in running the code at different degrees of throttle valve opening, from minimum to WOT, and for each throttle position at different rotational speeds, from minimum to maximum, by 250 rpm increases.

The Boost program"s output data were then post-processed to obtain maps of engine fuel consumption, f m  , and of the heat transfer rate between combustion gases and engine walls, gw Q  , as a function of brake mean effective pressure (BMEP) and engine rotational speed. The latter information on the thermal behaviour of the engine would have been particularly difficult to obtain via an experimental approach.

As examples of the results, figures 2 and 3 show the engine rotational speed-MEP charts and the iso- At any given working point, with the engine cold, fuel consumption is significantly greater (figure 2) and the combustion gases transfer to the walls a larger amount of heat per unit time (figure 3).

Cooling system model

This section deals with the thermo-fluid-dynamic models for internal and external engine cooling systems, which are described in sections 3.2.1 and 3.2.2, respectively. The coolant is assumed to exchange heat only in the radiator and the engine; all the other components are assumed to be adiabatic in the model.

Internal cooling system

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The internal cooling system consists of the coolant paths in the engine. Three main components were considered as parts of the internal cooling system: the liner, the coolant and the engine block (see figure 4). Accordingly, three enthalpy balances were formulated using a lumped parameter approach: Enthalpy balance for the liner:

dt dT c m Q Q w l l wc gw       (1)
Enthalpy balance for the coolant:

dt dT c m Q T c m T c m Q out e c c e c eb out e c c c in e c c c wc _ , , _ , _ ,               (2)
Enthalpy balance for the engine block, whose temperature is assumed to be the same as the oil"s:

dt dT c m c m Q Q Q eb oil oil eb eb a f eb        ) (    (3)
The heat transfer rates from liner walls to coolant, wc Q  , from coolant to engine block, eb Q  , and from coolant to ambient air, a Q  , were calculated using the following relationships:

             2 _ _ in c out c w l wc wc T T T A Q   (4)              eb in c out c eb eb eb T T T A Q 2 _ _   (5) ) ( a eb a a a T T A Q       (6)
The coolant to engine block heat transfer coefficient, eb  , was obtained using the Dittus-Boelter correlation (eq. A. 1 in the appendix) [START_REF] Kreith | Principles of Heat Transfer[END_REF].

The engine block to air heat transfer coefficient was calculated using the correlation [START_REF] Kreith | Principles of Heat Transfer[END_REF]:

           a eb a L   8 . 0 3 / 1
Re Pr 036 . 0 [START_REF] Guzzella | Introduction to Modeling and Control of Internal Combustion Engine Systems[END_REF] Prandtl"s and Reynolds" numbers in equation ( 7) regard the air flowing around the engine block, calculated for ambient conditions; its velocity was calculated by reducing vehicle speed by a factor that accounted for engine compartment geometry.

Evaluation of the heat transfer coefficient between engine walls and coolant, wc  , is complicated by the fact that at high engine loads the engine wall temperature in some regions may be higher than the coolant boiling temperature. In such conditions nucleate boiling may occur, and use of correlations that take into account only liquids would result in significant errors.

The details of the procedure used to calculate wc  are reported below in section 3.2.1.1.

  e f V FMEP Q     (8)
FMEP in [START_REF] Tumus | e-Thermal: A Vehicle-Level HVAC/PTC Simulation tool[END_REF] was evaluated by means of the ETH friction model relationship [START_REF] Guzzella | Introduction to Modeling and Control of Internal Combustion Engine Systems[END_REF]:

B k S k k T k T FMEP e e 4 2 2 3 2 1 ) ( ) ( ) (             (9)
In [START_REF] Luptowski | Development of the Enhanced Vehicle and Engine Cooling System Simulation and Application to Active Cooling Control[END_REF] the engine thermal condition is taken into account through term ) (

) ( ) ( 1 1   T k T k T e e  ,
          m k m s k k Pa T k 3.2.1.1 Wall to coolant heat transfer coefficient M A N U S C R I P T A C C E P T E D ACCEPTED MANUSCRIPT
As noted above, evaluation of the heat transfer coefficient between engine walls and coolant , wc  , is complicated by the fact that at high engine loads the engine wall temperature may exceed the coolant boiling temperature in some regions, possibly resulting in nucleate boiling. Use of correlations that consider only liquids would thus involve significant errors [START_REF] Kroes | Modeling of nucleate boiling in engine cylinder head cooling ducts[END_REF]. To take this into account, the effects of nucleate boiling were integrated in the model using Kandlikar"s correlation [START_REF] Kandlikar | A general correlation for saturated two-phase flow boiling heat transfer inside horizontal and vertical tubes[END_REF]. The correlation was used in the regions where the liner temperature exceeded the coolant boiling temperature, whereas the classical Dittus-Boelter correlation was used elsewhere. In this case the temperature distribution in the liner was assumed to be a 3 rd order polynomial (eq. 10) of the normalized height of the liner,

h , whose coefficients 1  , 2
 and 3  were fitted to the experimental data (from [START_REF] Heywood | Internal combustion engine fundamentals[END_REF]) of an engine running in fully warmed up condition (figure 6):

   h γ h γ h γ (0) T ) (h T (0) T (h) T 3 2 2 3 1 l l l l l         (10) l h h   0
In equation ( 16) ) 0 ( Once the temperature distribution in the liner is known, it is possible to determine the height at which the coolant boiling temperature is reached, eboil h (shown graphically in figure 4). The zone of the liner above

External cooling system

The external cooling system consists of the pump, radiator, thermostatic valve and external connection pipes.

The coolant mass flow c m  was obtained from the characteristic curves of the pump and the cooling circuit.

The thermostatic valve behaves as a proportional temperature controller, with a non-linear static gain; the valve time constant is approximately 30 seconds [START_REF] Guzzella | Introduction to Modeling and Control of Internal Combustion Engine Systems[END_REF]. As the coolant temperature at the engine outlet The coolant mass flow through the radiator at a given temperature was assumed to be proportional to valve lift and was thus calculated as:

c out e c t r m T l m     ) ( _ , ( 12
)
In a simplified circuit (figure 9) without a cabin heater, EGR cooler or oil cooler, the by-passed coolant mass flow becomes: [START_REF] Kroes | Modeling of nucleate boiling in engine cylinder head cooling ducts[END_REF] The enthalpy balance for the mass of coolant in the radiator is:

c out e c t by m T l m      )] ( 1 [ _ ,
dt dT c m T c m Q T c m out r c c r c out r c c r r out e c c r _ , , _ , _ ,             (14)
The coolant to air heat transfer rate was calculated by:

M A N U S C R I P T A C C E P T E D ACCEPTED MANUSCRIPT              a in e c out e c r r r T T T A Q 2 _ , _ ,   (15)
For any given radiator arrangement, the heat transfer coefficient, r  , increases with air velocity, hence with vehicle speed, a v , which can be calculated from the engine angular speed,  e , the tyre rolling circumference,  t , and the engine to wheel transmission ratio,  : [START_REF] Ferrari | Motori a Combustione interna[END_REF] Tables from [START_REF] Ferrari | Motori a Combustione interna[END_REF] allow a good estimation of r  as a function of a v for typical automotive applications.

  2π ω Γ v e t a

Pump inlet

The mass flow of coolant from the by-pass circuit and the mass flow of coolant from the radiator mix at the pump inlet.

Since all components except the engine and the radiator are assumed to be adiabatic, the coolant temperature in the pump does not change, and the same temperature is found both at the pump outlet and at the engine inlet.

The latter can therefore be calculated by solving the enthalpy balance at the pump inlet:

c c r_out , c c r e_out , c c by e_in , c c m T c m T c m T           (17)

DETERMINATION OF VEHICLE FUEL CONSUMPTION

After simulating the engine and its cooling system using the approach described in the previous section, we simulated a vehicle equipped with them running on standard cycles. Then, simulations of the same vehicle equipped with the perfect cooling system were run to identify its benefits in terms of fuel consumption.

The block diagram of the simulator used is shown in figure 10. The specifications of a high-medium class passenger car were entered: mass, 1500 kg; frontal area, 2.6 m 2 , drag resistance coefficient, 2.8, and a five-speed gearbox.

The engine torque and rotational speed required to run the test cycles were calculated and entered into the engine maps to obtain for each instant:

 the heat transfer rate between combustion chamber and liner

 engine-specific fuel consumption

Since these depend on the engine thermal conditions, the external engine cooling system model was also solved by using equations 1 to 17 to obtain the relevant temperatures of engine components, coolant and oil.

Knowledge of the fuel consumption values in the selected cycle and of the distance travelled allowed calculation of the amount of fuel needed to travel 100 km.

RESULTS AND DISCUSSION

This section reports a selection of the results obtained using the proposed simulator.

First of all the vehicle"s behaviour with the conventional cooling system is described by the aid of figure 11, which shows vehicle velocity, liner temperature, coolant temperature at the engine inlet and outlet, oil temperature, thermostatic valve position and engine fuel consumption for the NEDC.

Since the cycle starts with a cold engine, the first step of the simulation involves only ambient temperatures. As the engine warms up the liner temperature rises and coolant temperatures at the engine outlet and inlet grow almost simultaneously. Then, when the engine outlet temperature has reached about 60 °C, the thermostatic valve begins to open, and some of the coolant flows through the radiator, where it is cooled. From this point on, the engine inlet temperature is lower than that at the engine outlet. The position of the thermostatic valve changes according to the behaviour of its wax temperature-sensitive part. The oil temperature, which in the model is assumed to be equal to that of the engine block, behaves like the liner and the coolant temperature, its mean value rising with cycle time but much more slowly, due to its greater mass and greater thermal inertia.

The effect of smart thermal management is then demonstrated by comparing the results from the conventional cooling system to those obtained using the "perfect" cooling system, defined in section 2. This is done in figure 12, where two different perfect cases are compared to the standard case using the NEDC. Case 1 is a vehicle equipped with a cooling system that can warm up the engine instantly and maintain a constant set point temperature (this system fulfils principles #1, #2 and #3 of section 2. Case 2 applies the more realistic principle #3a instead of #3: the engine is adiabatic during the warm-up phase, i.e. the heat fluxes to the ambient through the engine block and the coolant are assumed to be zero; when the temperature set point is reached the cooling system converges with the one modelled in case 1.

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As shown in figure 12, the liner temperature set point, which is assumed to be the maximum allowable mean liner temperature, is reached immediately in case 1, it is reached in about 300 seconds in case 2, and is not reached at all with the conventional system.

The delay in reaching the fully warmed up condition and/or the imperfect control of the engine thermal status involve significantly greater friction losses, as shown in figure 13, where the FMEP (calculated by eq. 9) for Case 2 and for the conventional cooling system are compared for the NEDC normalized cycle. Figure 14 shows the heat generated per unit time that is associated with these friction losses.

Figure 14 compares the conventional and the perfect cooling system (Case 2) in terms of heat flow rate transferred from the in cylinder gases to the walls during the NEDC cycle; with the conventional cooling system more heat than necessary is transferred to the walls, increasing the thermal losses.

Figure 14 also shows the heat transferred to the ambient per unit time by the engine with the conventional cooling system, something that in the perfect cooling system is prevented from happening.

The conventional cooling system involves significantly greater fuel consumption, as shown in table 2, where Case 1, Case 2, and the conventional cooling system are compared for the NEDC, the EU CITY, and the most common US and Japanese normalized cycles.

Savings are greatest in the cycles simulating urban conditions, where engine loads are lowest. In such conditions an improved thermal management system offers the greatest advantages, whereas a conventional system is associated with mean engine temperatures that are well under fully warmed up values, excessively low oil temperature, and greater friction.

CONCLUSIONS

A method to evaluate the effects of an improved engine thermal management system for a passenger car is proposed.

A "perfect" cooling system was simulated and its effects on the engine thermal status defined with a view to establishing a reference to which different cooling and control system concepts could be compared.

A conventional cooling system was then compared to the "perfect" cooling system to estimate the fuel economy that can be achieved through the adoption of thermal management in a passenger car run on standard driving cycles.

Two different cases were hypothesized for the "perfect" cooling system: a system that warms up the engine instantly and maintains its set point temperature constant, and a more realistic case where the engine is assumed to be adiabatic while warming up. When the temperature set point is reached the two cases converge, because the cooling system is again assumed to be able to maintain its set point temperature constant.

The data required for the engine thermal modelling were obtained from a preliminary analysis using the AVL Boost program. Models for all the components of the power train and cooling system were developed and linked to simulate an upper-medium segment passenger car, which was then run on standard cycles. The use of a perfect system means that the simulated fuel consumption reductions constitute theoretical maxima for this car.

The results showed that delays in reaching the fully warmed-up condition and/or imperfect control of the engine thermal status both influence fuel consumption. Greater advantages were found in the cycles simulating urban conditions, where engine loads are lowest. These are the conditions in which an improved thermal management system offers the greatest gains, whereas in a conventional cooling system the mean engine temperatures are well under fully warmed up values, oil temperature is excessively low, and friction is greater.

Possible directions for research continuation include the collection of experimental data to verify the results of the simulation as well as result generalization by extending the data to different classes of cars.

APPENDIX A. Dittus-Boelter correlation

The Dittus-Boelter correlation [START_REF] Kreith | Principles of Heat Transfer[END_REF] calculates the heat transfer coefficient in horizontal tubes, c  , as: The values of constants 1 C -5 C are reported in table 3.

           
In any given condition the heat transfer coefficient, tp  , is obtained using the two sets of constants for the two regions.

Since the transition from one region to the other occurs at the intersection of the respective correlations, the higher transfer coefficient value is the predicted value of the proposed correlation.

The Kandlikar correlation is used in the paper to calculate the coolant to wall heat transfer flux in the region where the engine liner walls reach temperatures that exceed the coolant boiling temperature.

The dryness fraction, x, is assumed to be proportional to the difference between mean liner temperature and coolant boiling temperature.

Figures and tables captions

  between engine block and air (m 2 ) eb A heat transfer area between coolant and engine block (m 2 ) l A heat transfer area between liner and coolant (m 2 ) r A radiator heat transfer area (external) (m 2 ) B cylinder bore (m) BMEP brake mean effective pressure (Pa) ,(bar) c c coolant specific heat capacity (J kg -1 K -1 ) eb c engine block specific heat capacity (J kg -1 K -1 ) l c liner specific heat capacity (J kg -1 K -1 ) oil c oil specific heat capacity (J kg -1 K -1 ) FMEP engine friction mean effective pressure (Pa) ,(bar) eboil h liner height at which liner temperature equals coolant boiling temperature (m) in the engine (kg) r c m , mass of coolant in the radiator (kg) eb m engine block mass (kg) f m  fuel consumption (g s -1 ) through the radiator (kg s -1 ) by m  coolant mass flow through the by-pass duct (kg s -1 ) c m  coolant mass flow through the engine (kg s -1 ) r m  coolant mass flow through the radiator (kg s -1 ) to air heat transfer rate in the radiator (W) wc Q  liner to coolant heat transfer rate (W) engine temperature (K) V engine displacement (m 3 ) a v air velocity through the radiator (m s -1 ) a  engine block to air heat transfer coefficient (W m -2 K -1 ) eb  coolant to engine block heat transfer coefficient (W m -2 K -1 ) r  coolant to air heat transfer coefficient (W m -2 K -1 ) wc  wall to coolant heat transfer coefficient (W m -2 K -1 )  engine revs per cycle (rad s -1 )

  f m  and iso- gw Q  curves for the cold and the fully warmed up engine. Figure 2 also reports the engine working points for the European City driving cycle (EU CITY) and the New European Driving Cycle (NEDC) (see section 4 for vehicle details).

fQ

   in (3) represents the heat transfer rate due to mechanical friction, and can be obtained from the engine Friction Mean Effective Pressure (FMEP):

  temperatures at the bottom and top of the liner, respectively. A generic engine thermal condition can be described by adapting the pairs of shown in figure7, which reports the temperature distributions calculated with equation[START_REF] Yoo | An Engine Coolant Temperature Model and Application for Cooling System Diagnosis[END_REF] for liner mean temperatures ranging from cold start to fully warmed up conditions.

  affected by nucleate boiling, whereas only forced convection should occur below it.The heat transfer coefficients for the two zones, 1  and 2  , were calculated using the Dittus-Boelter and the Kandlikar correlations, respectively (eqs. A. 1 and B. 1 in the appendix), and then surface-weighted to obtain the mean heat transfer coefficient for the whole liner wc

  thermostatic valve lets more coolant through the radiator, thus increasing r m  and simultaneously reducing the coolant mass flow by m  in the by-pass circuit.

2 

 2 the coolant (W m -1 K -1 ) D e equivalent diameter of the engine coolant path around the liner (m) In the paper this correlation is used to calculate the coolant to wall heat transfer flux. The Reynolds number, Re, and Prandtl"s number, Pr, of the coolant are calculated based on cooling pump flow rate, cooling path section around the liner, and coolant mean temperature. B. Kandlikar correlation M A N U S C R I P T A C C E P T E D ACCEPTED MANUSCRIPT This correlation has been proposed to evaluate the heat transfer coefficient in horizontal tubes, tp  Froude number (-) g gravity acceleration (m s -2 ) The value of Co divides the whole surface in two regions: Co<0.65 convective boiling region (convective mechanism) Co>0.65 nucleate boiling region (nucleate boiling mechanism)

FMEP perfect cooling system case 2 FMEPFIGURESFigure 1 -Figure 2 -Figure 3 -Figure 4 -Figure 5 TFigure 6 -Figure 7 -TFigure 8 -Figure 9 -Figure 10 -Figure 11 -Figure 12 -Figure 13 -Figure 14 -

 21234567891011121314 Figure1Click here to download high resolution image

  whose diagram is reported in figure5as a function of the difference between maximum engine temperature,  T , and actual engine temperature,

																												T .
																												e
	The values for the coefficients used in (9) are:
	1	(	)	3	10	5	[	]	;	2	6 . 0	;	3	6 . 2	10	4	[	2	/	2	]	;	4	029 . 0	[	]	;	1

Table 1 -

 1 TABLES Engine characteristics

Table 2 -

 2 Comparison of fuel consumption data

Table 3 -

 3 Kandlikar's coefficients; C 5 = 0 for horizontal tubes with Fr > 0.04

ACKNOWLEDGEMENTS

The authors are grateful to Dr. Silvia Modena for the language review

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