A national laboratory of the U.S. Department of Energy

Office of Energy Efficiency & Renewable Energy

National Renewable Energy Laboratory 

Innovation for Our Energy Future 

Single-Phase Self-Oscillating 

Jets for Enhanced Heat Transfer 

Preprint 

S. Narumanchi, K. Kelly, and M. Mihalic 

National Renewable Energy Laboratory 

S. Gopalan and R. Hester 

Bowles Fluidics Corporation 

A. Vlahinos 

Advanced Engineering Solutions, LLC 

Presented at SemiTherm 2008 

San Jose, California 

March 16-20, 2008 

Conference Paper 

NREL/CP-540-42350 

June 2008 

NREL is operated by Midwest Research Institute ● Battelle Contract No. DE-AC36-99-GO10337 
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 1 

Single-Phase Self-Oscillating Jets for Enhanced Heat Transfer 

Sreekant Narumanchi,* Kenneth Kelly, and Mark Mihalic 

National Renewable Energy Laboratory 

1617 Cole Boulevard, Golden, CO 80401-3393 

*Email: Sreekant_narumanchi@nrel.gov 

Phone: 303-275-4062 

Shridhar Gopalan and Russ Hester 

Bowles Fluidics Corporation 

6625 Dobbin Road, Columbia, MD 21045 

Andreas Vlahinos 

Advanced Engineering Solutions, LLC 

4547 North Lariat Drive, Castle Rock, CO 80104 

Abstract 

In hybrid electric vehicles (HEVs), the inverter is a critical 

component in the power module, which conditions the flow of 

electric power between the AC motor and the DC battery 

pack. The inverter includes a number of insulated gate bipolar 

transistors (IGBTs), which are high-frequency switches used 

in bi-directional DC-AC conversion. The heat generated in the 

IGBTs can result in degraded performance, reduced lifetime, 

and component failures. Heat fluxes as high as 250 W/cm2 

may occur, which makes the thermal management problem 

quite important. In this paper, the potential of self-oscillating 

jets to cool IGBTs in HEV power modules is investigated 

experimentally. 

A full factorial design of experiments was used to explore 

the impact of nozzle design, oscillation frequency, jet flow 

rate, nozzle-to-target distance, and jet configuration (freesurface or submerged) on heat transfer from a simulated 

electronic chip surface. In the free-surface configuration, selfoscillating jets yielded up to 18% enhancement in heat transfer 

over a steady jet with the same parasitic power consumption. 

An enhancement of up to 30% for the same flow rate (and 

velocity since all nozzles have the same exit area) was 

measured. However, in the submerged configuration, amongst 

the nozzle designs tested, the self-oscillating jets did not yield 

any enhancements in heat transfer over comparable steady 

jets. Results also suggest that oscillating jets provide a more 

uniform surface temperature distribution than steady jets. 

Keywords 

Self-oscillating, jets, IGBTs, power electronics, freesurface, submerged, heat transfer, cooling 

Nomenclature 

A Area of target surface, m2 

AC Alternating current 

APEEM Advanced Power Electronics and Electrical 

Machines 

BFC Bowles Fluidics Corporation 

DC Direct current 

DOE Department of Energy 

h Heat transfer coefficient, W/m2

K 

HEV Hybrid electric vehicle 

IGBT Insulated gate bipolar transistor 

lpm Liters per minute 

NIST National Institute of Standards and Technology 

NREL National Renewable Energy Laboratory 

q Heat dissipated in the film heater, W 

T Temperature, K 

Subscripts 

ref Reference 

1. Introduction 

The overall goal of the FreedomCAR and Fuel Partnership 

is to reduce petroleum consumption by developing 

technologies that enable cars and trucks to become highly 

efficient, through improved power technologies and cleaner 

domestic fuels. To be effective, these technologies must be 

competitive in terms of cost and performance [1]. 

An electric propulsion system is essential for near-term 

combustion engine HEVs and longer-term fuel cell hybrid 

vehicles. Electric propulsion systems currently are available 

in the market but at a cost of approximately $33/kW [2]. The 

Department of Energy (DOE) Advanced Power Electronics 

and Electrical Machines (APEEM) activity seeks to reduce 

this cost to $12/kW by 2015 and to $8/kW peak by 2020, 

thereby enabling large-scale market penetration of advanced, 

energy-efficient vehicles [3]. The development of low cost, 

effective thermal management for power electronics and 

electrical machines is critical for the performance and 

reliability of the electronics system and for meeting the 

aggressive program cost targets [4]. 

The National Renewable Energy Laboratory (NREL) leads 

research and development activities in thermal control related 

to the DOE’s APEEM program. The overall objective of these 
thermal control activities is to develop advanced technologies 

and effective integrated thermal control systems aimed at 

meeting the FreedomCAR program goals for power module 

volume, cost, and weight. 

High heat-transfer rates are required to enable reductions 

in the cost, weight, and volume of power electronics 

components. Single-phase jets have been studied very 

extensively in the literature in the context of electronics 

cooling [5-8]. Some studies in the literature demonstrated that 

oscillating jets could enhance heat transfer over steady jets [9-

13]. In some of these studies, the oscillation was induced by 

mechanical means [9-12], while in others [13] the jets were 

self-oscillating. Self-oscillating jets involve no moving parts, 

and the oscillations are induced strictly by pressure differences 

in the fluid. This makes them quite attractive. However, there 

is very limited literature on the heat-transfer performance of 

self-oscillating jets. NREL established a collaboration with 

Bowles Fluidics Corporation (BFC) to explore the heattransfer performance of different types of fluidic nozzles. To 

conduct this study, NREL built a test fixture to accurately 

characterize the heat-transfer performance of different types of 

fluidic nozzles. Several fluidic nozzles were provided by BFC, 

and heat-transfer experiments were performed (with deionized 

water) to characterize the performance of these nozzles. 

Comparable experiments with non-oscillating, steady jets were 

also performed to serve as a baseline against which to 

compare the performance of the fluidic nozzles. 

2

Figure 1: Different jet impingement configurations (taken 

from [14]): (a) free-surface jet, (b) submerged jet, (c) confined 

submerged jet.

2. Description of a fluidic nozzle and self-oscillating jet 

 In the literature, the three main categories of jets that have 

been explored are free-surface, submerged, and confined 

submerged jets. These are illustrated schematically in Figure 1 

(taken from [14]). In a free-surface jet configuration [Figure 

1(a)], the liquid jet flows into another fluid—usually a gas— 

which is of much lower density than the liquid. In a 

submerged jet, the liquid jet flows into another fluid of 

comparable density [Figure 1(b)]. In a confined submerged jet 

[Figure 1(c)], the liquid jet is submerged, but there is also a 

plate or wall confining the liquid after impingement on the 

target surface. 

Figure 2: Illustration of the principle behind self-oscillating 

jets, (a) cross-sectional view of fluidic nozzle, (b) different 

nozzles tested. 

A self-oscillating jet is generated by a fluidic nozzle. 

Figure 2(a) illustrates the concept behind self-oscillating jets. 

However, the fluidic geometry illustrated is not an exact 

representation of the nozzle designs that were tested in this 

study. The oscillations are caused by the geometry inside the 

nozzle. The nozzle geometry causes vortices in the fluid, and 

the interacting vortices and consequent pressure differences in 

the fluid result in oscillations in the liquid jet at the nozzle 

exit. The jet oscillates within the angle depicted in Figure 2(a). 

This angle is referred to as the fan angle. The frequency of 

these oscillations is determined by the time taken for the jet to 

go from A to B and then back to A. Essentially, the jet 

performs a sweeping motion within the fan angle. 

In this study, we looked at nozzles that generate selfoscillating jets with different frequencies and fan angles. They 

are called self-oscillating because no moving parts are 

required to generate the oscillations. Typically, both the 

frequency and fan angle are a function of flow rate. Figure 

2(b) shows all the different fluidics encased in a housing, and 

gives a sense of the dimensions. For the particular nozzles 

used in this study, the oscillations are in a single plane. However, this need not always be the case. Nozzles can be 

designed to yield oscillations in the out-of-plane direction. 

A B 

Fan angle 

Jet Target 

surface

(a) 

(b) 

(b) (a)

(c)
3

Further details regarding the nozzles will be discussed in the 

next section. From a heat-transfer viewpoint, this sweeping 

motion has the potential to disrupt the boundary-layer growth 

on the target surface, and thereby enhance heat transfer 

between the surface and the liquid. 

Top Teflon 

block

Polycarbonate 

cylinder 

Pipe 

Reservoir 

Pump 

Filte

Valve

r

In-line 

heater 

Mass 

flow 

meter 

Turbine 

flow 

meter 

Valve 

Differential 

pressure 

transducer 

Exit/ 

Drain

Lower Teflon 

block 

Nozzle 

DC 

power supply 

Heater 

Insulation 

Copper 

block 

Thermocouples 

Liquid jet 

(a) 

(b) 

Top Teflon 

Pressure relief 

block 

Bottom Teflon 

block 

Drain 

Polycarbonate 

cylinder 

Well 

Nozzle 

Pipe 

Recess for 

copper plug 

3. Experimental loop and test fixture 

NREL built a high-heat-flux test loop to conduct 

experiments with single-phase heat exchangers. A schematic 

of the loop is shown in Figure 3(a). It consists of a reservoir, 

pump, filter, mass flow meters, turbine flow meters, and 

valves for regulating flow rate. The loop is also fully 

instrumented with pressure sensors for measuring fluid 

pressure drop, thermocouples for monitoring temperatures, 

and a data acquisition system controlled by LabView software. 

NREL also fabricated a test fixture to be used with the 

high-heat-flux test loop to characterize the heat-transfer 

performance of single-phase jets. The apparatus is shown in 

Figure 3(b). It consists of two Teflon plates with a transparent 

polycarbonate cylinder in between. A copper plug 3.8 cm 

(1.5”) in length and 1.27 cm (0.5”) in diameter fits snugly in 

the center of the lower Teflon plate. Teflon serves as an 

insulation to prevent heat loss and ensure one-dimensional 

heat flow in the copper plug. The surface of the copper plug 

Figure 3: (a) Schematic of test loop, (b) experimental apparatus. 
serves as the target surface on which the jet impinges [Figures 

4(a), 4(b)]. We conducted experiments with six different 

fluidic nozzles. The area of the jet exit of the fluidic nozzles is 

1.1 mm x 1.1 mm. The steady-jet nozzle has the same exit area 

of 1.21 mm2

 and has a circular opening of 1.24 mm diameter. 

For the steady jet, a circular opening was chosen to make the 

steady-jet nozzle yield a low pressure drop. The rationale was 

that this would provide a fair basis for comparison with the 

performance of the fluidic nozzles. 

4

Figure 4: Surface of the copper plug that serves as the target 

surface (a) top view, (b) cross-sectional view along the AB 

axis. 

The surface of the copper plug is shown in Figure 4(a). 

The jet centerline is aligned with the center point of the target. 

For self-oscillating jets, the plane of oscillation is also shown 

in the figure. There are four thermocouples on the surface of 

the copper plug. The rationale for placing multiple 

thermocouples was to determine the temperature distribution 

on the target surface. A resistive heater was attached to the 

back of the copper plug. Thermal grease was placed between 

the copper plug and the resistive heater to ensure good 

contact. A thermocouple was also attached to the heater to 

monitor its temperature. The power was supplied to the heater 

by a DC power supply. The maximum operating temperature 

of the heater is rated at 150°C. Key measurements included 

the coolant inlet temperature, the coolant outlet temperature, 

temperatures at four points on the copper target surface, the 

temperature of the resistive heater, the coolant flow rate, the 

pressure drop across the jet, and the power input to the heater. 

Thermocouples A numerical heat-transfer analysis was done (in FLUENT) 

on the lower Teflon plate and copper plug to estimate the heat 

loss from the sides and the back of the apparatus when heat is 

applied to the resistor. For this analysis, a heat-transfer 

coefficient was applied to the top of the copper target surface 

and the surrounding Teflon block. A wide range of heattransfer coefficients characteristic of impinging jets were 

applied on the copper surface (ranging from 20,000 to 50,000 

W/m2

K). From the simulations, we projected that the heat 

losses from the sides and back of the apparatus should be no 

more than 3%. The apparatus was placed on an insulating 

material with a very low conductivity of 0.02 W/mK. 

4. Experimental Results and Discussion 

We used a three-level (for velocity and target distance) 

full-factorial design-of-experiments to evaluate three jet 

velocities (2, 7, and 12 m/s) and three target distances (1.1, 

4.4, and 7.7 mm) in the free-surface and the submerged 

configurations (Table 1). Overall, we conducted 126 

experiments. This does not include repetitions that were done 

for some select experiments to establish the repeatability of 

these experiments. 

Using the experimental data, response surface functions 

for several performance attributes, such as temperature and 

heat transfer coefficient, were generated as a function of flow 

rate, target distance, nozzle type, and jet configuration. Three 

meta models were used — full 2nd-order polynomials, 

Kriging, and Neural Network. The quality of the response 

surface functions of all meta models was evaluated based on 

“goodness of the fit,” and the best response surface function 

was used in data post-processing. 

Table 1: Design of experiment variables. 

Nozzle design Configuration Target 

distance 

(mm)

Flow 

rate 

(lpm)

Steady Free Surface 1.1 0.15 

Fluidic 1 

Fluidic 2 4.4 0.5 

Fluidic 3 

Fluidic 4 Submerged 7.7 0.87 

Fluidic 5 

Fluidic 6 

Copper target surface, 

1.27 cm (0.5”) diameter 

Teflon 

(a) 

Jet plane 

A B

Teflon 

Film 

heater 

Copper (b) 

Insulation 
5

qu

all f

 with 1.1 

mm arget distance, (a) free-surface, (b) submerged. 

ted in this study, the heat-transfer 

coefficient is defined as 

As mentioned previously, the oscillating jets have an 

opening dimension of 1.1 mm × 1.1 mm. The baseline steady 

jet has a diameter of 1.24 mm and the same area (1.21 mm2

) 

as the fluidic nozzles. The pressure drop versus jet velocity, 

the fan angle versus velocity, and the frequency versus the 

velocity characteristics for the nozzles are shown in Figure 5. 

The behavior of the nozzles is quite diverse. For any given 

velocity/flow rate [Figure 5(a)], fluidic 1 has the highest 

pressure drop. As expected, the steady jet nozzle yields the 

lowest pressure drop. The fan angle versus velocity 

characteristics [Figure 5(b)] vary widely for the different 

Figure 5: Characteristics of the fluidic nozzles in the freesurface configuration, (a) pressure drop versus velocity, (b) 

fan angle versus velocity, (c) frequency versus velocity. 

 nozzles. Interestingly, fluidics 2 and 3 have a nearly constant 

fan angle over a wide velocity/flow rate range. Fluidics 5 and 

6 are closer to a steady jet in the sense that they have a very 

small fan angle; and because they are unstable, they do not 

have a well-defined frequency [Figure 5(c)]. For fluidics 1 

through 4, the frequency of oscillation increases as a function 

of flow rate/velocity [Figure 5(c)]. It is worth noting that 

fluidic 2 has considerably lower frequency of oscillation than 

the other nozzles. It is important to note that Figure 5 shows 

the characteristics of the fluidic nozzles in the free-surface 

[Figure 1(a)] configuration. At this point, the characteristics of 

the nozzles in the submerged configuration remain to be 

a

ad the lowest pressure drop at 

low rates for all nozzles. 

ntified. 

Figure 6 shows the pressure-drop versus flow-rate curves 

for the nozzles in the free-surface and submerged 

configurations at a target distance of 1.1 mm. Target distance 

did not have a significant impact on the pressure. Results with 

other target distances (4.4 and 7.7 mm) look similar to those 

presented in Figure 6. Configuration (free-surface or 

submerged) also did not affect the pressure drop. At any given 

flow rate, fluidic 1 had the highest pressure drop. At most flow 

rates or velocities (Figure 5), the frequency and fan angle of 

fluidic 1 were also the highest of the nozzles tested. Figure 6 

also shows that the steady jet h

(a) 

0

2

4

6

8

10

12

14

0 2 4 6 8 10 12

Velocity (m/s)

Pressure drop (psi)

fluidic1

fluidic2

fluidic3

fluidic4

fluidic5

fluidic6

0

5

10

15

20

25

30

0 0.2 0.4 0.6 0.8 1

Flow rate (lpm)

Pressure drop (psi)

steady

fluidic1

fluidic2

fluidic3

fluidic4

fluidic5

fluidic6

0

5

10

15

20

25

30

0 0.2 0.4 0.6 0.8 1

Figure 6: Pressure drop as a function of flow rate

 t

In the results presen

− ref )TT(A

where h is the heat-transfer coefficient, q is the power applied 

to the resistive heater, A is the area of the surface of the 

copper plug (1.267 cm2

), T is the surface temperature of the 

q h =

copper surface (average of the four thermocouple readings on 

Flow rate (lpm)

Pressure drop (psi)

steady

fluidic1

fluidic2

fluidic3

fluidic4

fluidic5

fluidic6

(a) 

(b) 

(b) 

0

10

20

30

40

50

60

0 2 4 6 8 10 12

Velocity (m/s)

Fan angle

fluidic 1

fluidic 2

fluidic 3

fluidic 4

fluidic 5

fluidic 6

(c) 

0

100

200

300

400

500

600

700

800

0 2 4 6 8 10 12

Velocity (m/s)

Frequency (Hz)

fluidic 1

fluidic 2

fluidic 3

fluidic 4
the copper target surface), and Tref is the jet inlet temperature. 

For the experiments reported here, Tref is 25°C. 

The experimental uncertainties in the heat transfer 

coefficients obtained in this study were determined by using 

the method given by Kline and McClintock [15]. The 

uncertainty levels presented are analogous to 95% coverage. 

The accuracy of the turbine flow meters was ±0.1% of the 

flow rate, while the repeatability was ±0.05% of the flow rate. 

The accuracy of the mass flow meter was ±0.2% of the flow 

rate, while the accuracy of the differential pressure transducer 

used to measure the pressure drop across the jet was ±0.5% of 

the full scale (0-30 psi). Temperature measurements were 

accurate to within 0.05 K. This includes any error due to the 

data acquisition system. All thermocouples were calibrated in 

a temperature bath. The accuracy of the temperature bath was 

confirmed to be within 0.03K of a NIST-traceable temperature 

Figure 7: Impact of target distance on heat transfer with 

fluidic 1 in the (a) free-surface, (b) submerged configuration. 

6

probe within the temperature range of interest (20°C to 50°C). 

Heat loss from the copper plug of approximately 3% was also 

factored into the analysis to yield an uncertainty in the heat 

transfer coefficient on the order of ±4%. Some select 

experiments were also performed to deduce the repeatability 

of the experiments. The repeatability of the heat transfer 

coefficients was within 2%. In the results presented later, 

parasitic power loss across the jet (product of pressure drop 

across the jet and the flow rate) is also a parameter. The 

uncertainty in the parasitic power was less than 0.6%. 

Figure 7 reveals the impact of target distance on heat 

transfer for fluidic 1 in both the free-surface and submerged 

configurations. For the free-surface configuration, at lower 

flow rates a larger target distance yielded a better heat-transfer 

performance than a shorter target distance. However, at higher 

flow rates, a shorter target distance yielded higher heattransfer performance than a larger target distance. Overall, in 

the free-surface configuration, the impact of target distance on 

heat transfer for the nozzles was noticeable, but not very 

strong. At low flow rates, the heat-transfer coefficients for the 

target distances differed by up to 27%; at elevated flow rates, 

the difference in the heat-transfer performance at different 

target distances was within 8%. 

However, in the submerged configuration, target distance 

had a big impact on the heat transfer performance of the 

different fluidic nozzles. As shown in Figure 7(b), as target 

distance increases, at any given flow rate the heat transfer 

attenuates. At elevated flow rates, the impact of target distance 

became even more pronounced. At the lowest flow rate 

[Figure 7(b)], the difference in heat transfer results for the 

different target distances was about 10%; for the highest flow 

rate, the difference was almost 35% to 40%. Though results 

are not shown here, similar trends were observed for other 

fluidic nozzles as well. This aspect was more pronounced for 

the fluidic nozzles than for the steady jets, because the 

oscillations and the sweeping motion are damped out with 

increasing target distance. 0

10000

20000

30000

40000

50000

60000

70000

0 0.2 0.4 0.6 0.8 1

Flow rate (lpm)

Heat transfer coefficient (W/m2K)

1.1 mm

4.4 mm

7.7 mm

Figure 7 reveals that at a short target distance (1.1 mm), 

the performance of fluidic 1 in the free-surface configuration 

was almost identical to that of the submerged configuration. 

However, at a greater target distance (7.7 mm), the freesurface configuration yielded much higher heat-transfer 

performance than the submerged configuration. The reason for 

this is most likely the damping effect mentioned previously. 

0

10000

20000

30000

40000

50000

60000

70000

0 0.2 0.4 0.6 0.8 1

Flow rate (lpm)

Heat transfer coefficient (W/m2K )

1.1 mm

4.4 mm

7.7 mm

(b) 

(a) 

Figure 8 shows the heat-transfer performance of different 

nozzles at the optimum target distance for the free-surface and 

submerged configurations. This optimum target distance for 

most cases was 1.1 mm. Figures 8(a) and 8(b) show the heattransfer coefficients as a function of flow rate and parasitic 

power, respectively, for the free-surface configuration. Figures 

8(c) and 8(d) show the same for the submerged configuration. 

For any given flow rate, different nozzles have different 

pressure drops, which means the parasitic power consumed 

(product of flow rate and pressure drop across the jet) is 

different. The performance of the different nozzles must be 

compared for the same parasitic power consumed. This is the 

rationale for plotting the heat transfer coefficient as a function 

of parasitic power. Figure 8(a) shows that, in the free-surface 

configuration, fluidic 1 yielded the highest heat-transfer 

coefficients at elevated flow rates (beyond 0.5 lpm). The same 

figure shows that the steady jet yielded the lowest heattransfer coefficient. This suggests that for a given flow rate, 

heat transfer was enhanced in the free-surface configuration 
with self-oscillating jets. The results also suggest that in the 

free-surface configuration, a higher frequency of oscillation 

leads to increased heat transfer. It is more illustrative to look 

at Figure 8(b), which shows the heat-transfer coefficient 

versus parasitic power for the free-surface configuration. 

Interestingly, this plot shows that fluidic 3 was the best 

performing nozzle. Fluidic 1 had a higher pressure drop than 

fluidic 3, so when its heat-transfer performance is plotted 

against the parasitic power, it did not rank the highest. Figure 

8(b) shows that the steady jet was the lowest-performing 

nozzle in the free-surface configuration. A hypothesis in the 

literature [16] indicates that at elevated flow rates in the freesurface configuration, splashing causes the fluid film to 

become very thin and causes bulk warming of the fluid 

because the boundary layer reaches the surface of the film and 

attenuates the heat-transfer coefficient. Also, at close target 

distances, the sweeping motion of the oscillating jet causes 

significant disruptions in the boundary layer growth, which 

reduces the fluid-to-solid heat-transfer resistance. Figure 8 

illustrates that, in the free-surface configuration, the fluidic 

nozzles can yield up to 30% enhancement over a steady jet at 

comparable flow rates. For the same parasitic power, the 

enhancement over a steady jet is about 18%. To get a sense of 

the numbers, at a flow rate of 0.87 lpm (velocity of 12 m/s), a 

heat-transfer coefficient of about 65,000 W/m2

K was achieved 

with fluidic 1. 

Figures 8(c) and 8(d) show heat transfer versus flow rate 

and heat transfer versus parasitic power, respectively, for the 

submerged configuration. Figure 8(c) shows that the curves 

for the nozzles are all linear. There are some differences in the 

heat transfer results from the nozzles. At low flow rates, the 

heat-transfer coefficients achieved with the different nozzles 

were within 20% of one another; at higher flow rates, they 

were within 8% of one another. However, Figure 8(d) shows 

that fluidic 5 and the steady jet are the best performing 

nozzles. In the submerged configuration, virtually no 

(a) 

7

(c) (d) 

0

10000

20000

30000

40000

50000

60000

70000

0 0.2 0.4 0.6 0.8 1

Flow rate (lpm)

Heat transfer coefficient (W/m2K)

fluidic 1

fluidic 2

fluidic3

fluidic 4

fluidic 5

fluidic 6

steady

0

10000

20000

30000

40000

50000

60000

70000

0 0.5 1 1.5 2 2.5 3

Parasitic power (W)

Heat transfer coefficient (W/m2K)

fluidic 1

fluidic 2

fluidic3

fluidic 4

fluidic 5

fluidic 6

steady

0

10000

20000

30000

40000

50000

60000

70000

0 0.2 0.4 0.6 0.8 1

Flow rate (lpm)

Heat transfer coefficient (W/m2K)

fluidic 1

fluidic 2

fluidic 3

fluidic 4

fluidic 5

fluidic 6

steady

0

10000

20000

30000

40000

50000

60000

70000

0 0.511.522.5 3

Parasitic power (W)

Heat transfer coefficient (W/m2K)

fluidic 1

fluidic 2

fluidic 3

fluidic 4

fluidic 5

fluidic 6

steady

(b) 

Figure 8: Heat-transfer performance of nozzles at optimum target distance: (a) heat-transfer coefficient (h) versus flow rate in 

the free-surface configuration, (b) h versus parasitic power in the free-surface configuration, (c) h versus flow rate in the 

submerged configuration, (d) h versus parasitic power in the submerged configuration. 
enhancement was achieved from the self-oscillating jets over a 

steady jet; i.e., the best performing fluidic nozzles are nearly 

identical to the steady jet. Another important point is that the 

best heat transfer at a given parasitic power in the submerged 

configuration was higher than that in the free-surface 

configuration. 

There may be two reasons for the results seen in Figure 8. 

First, in the submerged configuration, the steady-jet heattransfer performance picks up because there is no splashing, 

which is a problem as mentioned earlier for the free-surface 

steady-jet configuration. Second, in the submerged 

configuration, the strength of the oscillations of the oscillating 

jets is somewhat dampened and attenuated. This is particularly 

true at larger target distances. 

8

Our experiments showed that the steady submerged jet 

outperformed the steady free-surface jet by almost 30% in 

certain cases. This suggests that eliminating splashing and 

warming of the fluid in the boundary layer for the steady jets 

is very important. The nozzle characteristics presented in 

Figure 5 (fan angle versus velocity and frequency versus 

velocity) are only for the free-surface configuration. These 

characteristics could change considerably in the submerged 

configuration. 

The oscillating jets also appeared to increase the target 

surface temperature uniformity with respect to steady jets. 

Uniformity of target surface temperature is important. In a 

practical IGBT package, uniform temperatures will lead to 

reduced thermal stresses. Figure 9 shows the distribution of 

temperature across the surface. The temperatures are derived 

by a smooth fit of the four thermocouple surface temperature 

measurements. The mean target surface temperature (dark 

vertical line) and the standard deviation in the target surface 

temperature are also shown for the steady jet and the 

oscillating jet generated by fluidic 3. Both the free-surface and 

submerged configurations are shown. The results are 

presented for a target distance of 4.4 mm and a velocity of 7 

m/s. The y-axes show the temperature distribution on the 

target surface. The standard deviation of temperatures for the 

steady jet [Figures 9(a) and 9(b)] is much higher than that of 

the fluidic 3 jets [Figures 9(c) and 9(d)]. This suggests that 

oscillating jets are yielding more uniform target surface 

temperatures in both the free-surface and submerged 

configurations. It is also worth noting that, for fluidic 3 in the 

submerged configuration [Figure 9(d)], even though the 

overall heat transfer is poor compared with that of the steady 

submerged jet [Figure 9(c)], the temperature uniformity is 

much better. One interesting fact is that the submerged fluidic 

3 jet yields the lowest standard deviation in temperature 

Figure 9: Temperature distribution on the target surface for (a) steady free-surface jet, (b) steady submerged jet, (c) free-surface 

(a) 

(b) 

(c) 

(d) 

jet from fluidic 3, and (d) submerged jet from fluidic 3, for a target distance of 4.4 mm, and jet velocity of 7 m/s. 
9

amongst the four cases examined in Figure 9. Further research 

is needed to determine whether these trends in surface 

temperature uniformity are consistent among all the cases. 

 The results presented here with self-oscillating jets are not 

completely exhaustive. Other nozzle designs, such as larger 

diameter nozzles and nozzles generating out-of-plane 

oscillations, which might be better performing than those 

presented in this paper, could be explored. 

5. Conclusions 

 This paper reports on one of the very few experimental 

studies exploring the heat-transfer potential of liquid selfoscillating jets. We explored the heat-transfer performance of 

several self-oscillating jet designs. In the free-surface 

configuration, self-oscillating jets showed enhancements over 

comparable steady jets of almost 18% with the same parasitic 

power loss. For a given flow rate (and hence velocity), without 

limiting parasitic power loss, free-surface self-oscillating jets 

enhanced heat transfer up to 30% over that of a steady jet. In 

the free-surface configuration, results hinted that a higher 

frequency of oscillation led to increased heat transfer. 

However, in the submerged configuration, amongst the fluidic 

nozzles explored, we did not observe any heat-transfer 

enhancements with self-oscillating jets over a steady jet. 

Results suggest that oscillations can also improve the 

temperature uniformity of a target surface with respect to a 

steady jet. The nozzle designs that were investigated in this 

study are not exhaustive, and further variations of nozzle 

design will be explored in the future. 

Acknowledgments 

The authors would like to acknowledge the support 

provided by Susan Rogers, Technology Development 

Manager, Advanced Power Electronics and Electric Machines, 

DOE FreedomCAR and Vehicle Technologies Program of the 

Office of Energy Efficiency and Renewable Energy. The 

authors would like to acknowledge Desikan Bharathan and 

Keith Gawlik for helpful discussions regarding the 

experimental apparatus. 

References 

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 vehiclesandfuels/about/fcvt_mission.html. 

2. Energy and Environmental Analysis, Inc., Technology 

and cost of the MY2007 Toyota Camry HEV – Final 

Report, Oak Ridge National Laboratory Subcontract 

report, ORNL/TM-2007/132, 2007. 

3. Electrical and Electronics Technical Team, “Electrical 

and Electronics Technical Team Roadmap,” Nov. 2006, 

www.eere.energy.gov/vehiclesandfuels/pdfs/program/eett

_roadmap.pdf, 2006. 

4. Kelly, K., et al., “An overview of thermal control 

technologies for cooling electric vehicle power 

electronics,” Proceedings of the EVS 23 Conference, 

Anaheim, CA, Dec. 2-5, 2007. 

5. Garimella, S.V., “Heat transfer and flow fields in 

confined jet impingement,” in Annual Review of Heat 

Transfer, pp. 413−494, 2000. 

6. Lienhard, J.H., “Liquid jet impingement,” in Annual 

Review of Heat Transfer, pp. 199-270, 1995. 

7. Webb, B.W., and Ma, C.-F., “Single-phase liquid jet 

impingement heat transfer,” Advances in Heat Transfer, 

26, pp. 105−217, 1995. 

8. Narumanchi, S.V.J., Amon, C.H., and Murthy, J.Y., 

“Influence of pulsating submerged liquid jets on chiplevel thermal phenomena,” ASME Journal of Electronic 

Packaging, 125, pp. 354−361, 2003. 

9. Zumbrunnen, D.A., and Aziz, M., “Convective heat 

transfer enhancement due to intermittency in an 

impinging jet,” ASME Journal of Heat Transfer, 115, pp. 

91−98, 1993. 

10. Mladin, E.C., and Zumbrunnen, D.A., “Nonlinear 

dynamics of laminar boundary layers in pulsatile 

stagnation flows,” Journal of Thermophysics and Heat 

Transfer, Vol. 8, No. 3, pp. 514−523, 1994. 

11. Mladin, E.C., and Zumbrunnen, D.A., “Local convective 

heat transfer to submerged pulsating jets,” International 

Journal of Heat and Mass Transfer, Vol. 40, No. 14, pp. 

3305−3321, 1997. 

12. Wang, Y., Yuan, G., Yoon, Y.-K., Allen, M.G., and 

Bidstrup, S.A., "Optimization of synthetic jet fluidic 

structures in printed wiring boards," ASME Journal of 

Electronic Packaging, 128, pp. 353-359, 2006. 

13. Camci, C., and Herr, F., “Forced convection heat transfer 

enhancement using a self-oscillating impinging planar 

jet,” ASME Journal of Heat Transfer, 124, pp. 770−782, 

2002. 

14. Wolf, D.H., Incropera, F.P., and Viskanta, R., “Jet 

impingement boiling," Advances in Heat Transfer, 23, pp. 

1−132, 1993. 

15. Kline, S.J., McClintock, F.A., “Describing uncertainties 

in single-sample experiments,” Mech. Eng. (Am. Soc. 

Mech. Eng.), 75, pp. 3-8, 1953. 

16. Womac, D.J., Ramadhyani, S., and Incropera, F.P., 

“Correlating equations for impingement cooling of small 

heat sources with single circular liquid jets,” ASME 

Journal of Heat Transfer, 115, pp. 106−115, 1993. 
F1147-E(09/2007) 

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Single-Phase Self-Oscillating Jets for Enhanced Heat Transfer: 

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S. Gopolan, R. Hester: Bowles Fluidics Corp. 

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14. ABSTRACT (Maximum 200 Words)

In hybrid electric vehicles, the inverter is a critical component in the power module, which conditions the flow of 

electric power between the AC motor and the DC battery pack. The inverter includes a number of insulated gate 

bipolar transistors (IGBTs), high-frequency switches used in bidirectional DC-AC conversion. The heat generated in 

IGBTs can degrade performance, reduce lifetime, and cause component failures. Heat fluxes as high as 250 W per 

square centimeter may occur, which makes the thermal management problem quite important. In this paper, the 

potential of self-oscillating jets to cool IGBTs in HEV power modules is investigated experimentally. 

15. SUBJECT TERMS

hybrid electric vehicles; inverters; IGBTs; heat transfer; oscillating jets; vehicle power electronics 

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