Review of Two-phase Electronics Cooling for Army Vehicle 

Applications 

by Darin Sharar, Nicholas R. Jankowski, and Brian Morgan 

ARL-TR-5323 September 2010 

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ARL-TR-5323 September 2010 

Review of Two-phase Electronics Cooling for Army Vehicle 

Applications 

Darin Sharar, Nicholas R. Jankowski, and Brian Morgan 

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14. ABSTRACT 

Implementation of many Army vehicle electronics technologies is becoming increasingly dependant on the ability to dissipate 

large amounts of waste heat. In the past, these technologies have been primarily cooled with active and passive air or 

single-phase heat sinks. While single-phase active and passive cooling methods are well-characterized, making system 

implementation predictable and dependable, these schemes are not providing the cooling performance necessary for future 

high-heat-flux electronic systems. By comparison, two-phase cooling has demonstrated the ability to provide net benefits in 

heat dissipation, size, weight, and overall power consumption necessary for future electronic systems. However, unlike 

single-phase cooling, two-phase cooling performance is difficult to estimate, making system integration unpredictable and 

often unreliable. Therefore, two-phase research and system integration have become major focal points for academic and 

professional organizations, alike, aimed towards resolving the uncertainties associated with two-phase cooling and developing 

the next generation of high-heat dissipation techniques. This report reviews the major two-phase cooling approaches in 

context of Army vehicular applications. 

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iii 

Contents 

List of Figures iv

List of Tables iv

1. Introduction 1

1.1 Army Electronics Cooling Needs ....................................................................................1

1.2 Army Vehicle Platform Considerations ..........................................................................2

2. Two-Phase Cooling Background 4

2.1 Spray Cooling ..................................................................................................................7

2.2 Jet-impingement Cooling ................................................................................................9

2.3 Microchannel Cooling ...................................................................................................10

2.4 Technology Comparison ...............................................................................................12

3. Two-phase Microchannel Flow Obstacles 13

3.1 Microchannel Two-phase Steady-state Flow Regimes .................................................13

3.2 Thermal Failure and Critical Heat Flux (CHF) .............................................................13

3.3 Channel Orientation ......................................................................................................17

3.4 Microchannel Two-phase Transient Effects ..................................................................17

3.5 Microchannel Flow Summary .......................................................................................18

4. Two-phase Macro/Microchannel Commercial Success 18

5. Conclusions and Recommendations 22

6. References 24

Appendix A. Typical Coolant Fluids and Associated Properties 33

List of Symbols, Abbreviations, and Acronyms 34 

Distribution List 35 
iv 

List of Figures 

Figure 1. Primary components of a two-phase cooling loop. .........................................................5

Figure 2. Heat transfer coefficients possible with natural convection, single-phase liquid 

forced convection and boiling for various fluids; from (28). ....................................................6

Figure 3. Primary components of a spray cooler. ...........................................................................7

Figure 4. Qualitative comparison for heat transfer coefficients of spray coolers and jet 

impingement coolers; from (28). ...............................................................................................8

Figure 5. Primary components of a jet-impingement cooler. .........................................................9

Figure 6. Primary components of a microchannel cooler. ............................................................11

Figure 7. Ideal microchannel 2-phase boiling flow regimes in (a) horizontal and (b) vertical 

orientations; from (60). ............................................................................................................14

Figure 8. Critical Heat Flux for Water @ 1atm. Point A is the onset of boiling and graph 

follows typical flow progression up to annular flow and CHF at Point C; from (21). ............16

Figure 9. Non-Ideal microchannel 2-phase boiling flow regimes in (a) horizontal and 

(b) vertical orientations; from (60). .........................................................................................17

Figure 10. Two-phase Parker Hannifin 450A IGBT cooling module; from (14). ........................19

Figure 11. Cooling Performance of the two-phase Parker Hannifin cooler compared to 

single-phase cooling techniques; from (14). ............................................................................21

List of Tables 

Table 1. Parker-Hannifin inverter test cases; from (14). ..............................................................20
1 

1. Introduction 

1.1 Army Electronics Cooling Needs 

The increasing development of electric platforms has been a growing trend across the separate 

branches of the U.S. Department of Defense (1–5). The platform proponents hope that replacing 

many of the traditional mechanical power and control systems with electronic equivalents will 

enable significant improvements by reducing weight, lowering fuel usage, increasing design 

flexibility, and enhancing overall functionality (6). In addition to work in the commercial sector, 

this trend is apparent in “drive by wire” and “fly by wire” system designs (7, 8). Specific to U.S. 

Army platform development, the past few iterations of Army vehicle programs have focused on 

increasing the use of electrical systems throughout the vehicle for many of the same reasons (6, 

8–11). Some of these programs have focused on new fully hybrid vehicles, while others have 

looked at improvements that can be made to legacy systems (8). Because the disparate electrical 

systems common to these programs need to draw power from a platform’s common electrical 

bus, they all require improved power conversion electronics to meet operational specifications. 

U.S. Army Tank Automotive Research, Design and Engineering Center (TARDEC) and the U.S. 

Army Research Laboratory (ARL) have been actively investigating hybrid electric vehicle 

architectures, power electronic subsystems, and components in support of this development 

strategy (6, 12). Aligned efforts aimed towards increasing the component power while 

simultaneously decreasing component size have led to improvements in power density. At the 

same time, however, increasing power density directly exacerbates the task of managing waste 

heat while staying within acceptable temperature limits for these components (9); new generation 

electronic systems are producing multiple kilowatts of waste heat with projected future device 

level heat fluxes upwards of 1000 W/cm2 (13). 

Most vehicle designs have relied on air-cooled heat sinks for low-heat-flux electronics and an 

antifreeze-based automotive liquid coolant loop to manage the larger waste heat from engine 

compartment components (11, 14). However, new power-dense electronic systems are further 

increasing waste heat and presenting great challenges to the capabilities of conventional air and 

single-phase liquid cooling systems (2). Considering strictly air cooling, the effect of higher heat 

flux electronics is larger, heavier, costlier heat sinks and fans to compensate for insufficient 

convective performance. The effect is equally dramatic with single-phase liquid cooling, with 

higher heat fluxes requiring larger coolant flow rates to sufficiently cool the system devices (15). 

These large flowrates and subsequent pumping powers result in increasingly bulky, heavy 

systems that consume more fuel (10). Thus, there is a drive to develop improved cooling 

components that are smaller and lighter, and have increased convective performance relative to 

conventional liquid cold plates. 
2 

Over the past several decades, the academic and industrial research communities have explored 

many options for improving liquid cold plate convective performance. These options have 

included implementing various micro- and mini-channel designs to increase heat transfer surface 

area, as well as using alternative fluid delivery methods in the forms of both macro- and microscale jets and sprays (16–18). No matter the particular heat sink configuration, the automotive 

cooling options that have been pursued beyond a laboratory environment have almost 

exclusively been single-phase liquid flow designs (7). Extensive single-phase research and 

design iterations have, until recently, provided sufficient thermal performance for cooling Army 

vehicles. 

With ever-increasing electronic heat fluxes and mounting single-phase thermal management 

issues, cooling schemes using liquid-vapor phase change (hereafter referred to as “two-phase 

cooling” in this report) have been examined as a practical and cost-conscious next-steps beyond 

single-phase cooling. In fact, the Department of Defense (DoD) and the National Aeronautics 

and Space Administration (NASA) have already distinguished two-phase flow technologies as a 

favorable solution to meeting the strict demands for some emerging cooling requirements (3, 4). 

Despite these endorsements, two-phase cooling schemes have traditionally not been used due to 

technical risks related to flow, pressure, and critical thermal failure. This report will review both 

the potential advantages and disadvantages of two-phase cooling technology within the context 

of Army vehicle power system thermal management, and propose paths for future research. 

1.2 Army Vehicle Platform Considerations 

We examined the Army vehicle platform and the constraints it imposes upon a particular cooling 

solution. As previously mentioned, most vehicles have a single-liquid coolant loop for the 

removal of waste heat from the engine and other “under hood” components. This has historically 

been an antifreeze loop consisting of a 50/50 ethylene glycol and water (EGW) solution (14, 19), 

though there has been a recent commercial movement toward the use of lower toxicity propylene 

glycol and water (PGW) solutions (20). This loop is driven by a single-coolant pump, and after 

acquiring heat from various heat sources will pass through a liquid-air heat exchanger (typically 

a fan-cooled radiator) to reject heat to the environment. 

The temperature at which the coolant is delivered to the electronic components is highly 

dependent on operating condition. Current Army vehicle requirements call out inlet 

temperatures around 65 °C, with future objectives of 80–100 °C (9) because higher coolant 

temperatures increase efficiency and improve power and weight savings on the radiator side of 

the cooling loop (21). Most silicon power electronics have a full-power operating temperature 

limit of about 125 °C (22), and the boiling temperatures for EGW and PGW at ambient pressures 

are 107 °C and 106 °C, respectively. These high inlet temperatures limit the electronic device 

temperature increase to only 25–45 °C above inlet temperature, reducing electronic cooling 

potential and requiring exceptional cooling performance to dissipate high-heat-fluxes and 

prevent overheating. 
3 

In addition to low allowable temperature margins, the EGW coolant, itself, limits the electronics 

thermal budget. EGW is used mainly to provide a wider liquid temperature range for the fluid 

than a water system would provide (–40 °C–107 °C compared to 0 °C–100 °C at ambient 

pressure, respectively). However, past research on single-phase cooling has shown that EGW 

and PGW mixtures can have a 28% to 42% reduction, respectively, in cooling performance 

compared to a pure water microchannel system (23), mainly due to increased viscosity and 

decreased specific heat of the EGW and PGW mixtures compared to pure water. Furthermore, 

typical cooling systems are not even pure 50/50 EGW. Additives that affect fluid properties—

borax, sodium Mercaptobenzothiazole (sodium MBT), trisodium phosphate, and antifoaming 

agents—are often present. It is unclear exactly what effect these additives have on thermal 

performance or the function of advanced heat exchangers (19). 

The single coolant loop architecture also restricts the potential performance of the electronic 

cooling components. Because the engine is usually the primary cooling concern for this loop, 

changes to loop parameters, such as fluid type, flowrate, and available pressure, generally cannot 

be made even if electronics cooling performance could be improved. The loop is generally high 

flowrate with low pressure drop available for electronics cooling (typically 2–5psi), which 

complicates or even prevents the use of many cooling solutions that often have driving pressure 

requirements on the order of 10–100 psi (16, 24, 25). While these designs may reduce total 

pumping power, integration to current single loop architectures would require flow conversion. 

This single loop architecture is also not designed to handle combined gas and liquid flow, as 

would occur in boiling heat transfer. Consequently, vapor buildup in any portion of the system 

could result in coolant venting or reduced coolant flow through the radiator and premature 

thermal failure due to subsequent reduced heat transfer. Therefore, a margin of safety is required 

to maintain the system well within single-phase conditions, which can increase flowrate, 

pumping power, and radiator requirements. Additionally, the specific boiling behavior of EGW 

and PGW (termed binary fluids) is typically suppressed compared to the onset conditions for 

boiling of pure components (26). The resulting suppression, whose severity is dependent on the 

thermodynamic properties of the specific mixture, presents an additional resistance to heat 

transfer. 

Finally, while the engine coolant loop is a sealed system, it is user-serviced and subject to 

contamination and particulate accumulation (19). From a design perspective, contamination 

either requires a system with large enough fluid lines and cooling geometries to provide 

immunity to clogging (27), or stringent filtering requirements to remove contamination from the 

system. Both of these scenarios have negative consequences; typically larger geometries in cold 

plates result in reduced thermal performance, and rigorous filtering may cause large system 

pressure drops and reduced flowrates. 

Because of the limitations of a single coolant loop, some consideration has been given to a 

secondary cooling loop that either exchanges heat with the primary coolant loop or with outside 
4 

air directly. This approach is most readily apparent in the cabin cooling or air conditioning 

systems present in almost all modern vehicles. Beyond this, other platforms have examined the 

use of secondary loops specifically designed for cooling non-engine electronic heat loads (3, 7). 

Four potential benefits lie in the ability to design a secondary cooling loop optimized for the 

specific heat loads in the system: 

1. Flexibility of working fluid, including refrigerants in some cases. Table 1 in appendix A 

shows typical coolant fluids with compiled properties. 

2. Possibility to implement a vapor compression refrigeration cycle to deliver coolant to the 

cold plates below ambient temperature. 

3. Pump autonomy to optimize system pressure, flow, and temperature to match specific heat 

loads. 

4. Potential for hermetically sealed systems to reduce contamination and clogging/fouling 

potential. 

There are definite disadvantages involved with implementing a secondary loop; the most obvious 

of these are the additional size, weight, and power (SWAP) required for the extra loop 

components such as pumps, hoses/tubing, and heat exchangers/condensers. While the possible 

drawbacks of such a two-loop system may be significant, the overall potential for cooling 

performance improvement may be able to offset those negative elements. Implementation of a 

secondary loop may add system complexity but enables the use of thermal techniques, like 

microchannels and two-phase cooling, that would be logistically difficult with a primary engine 

coolant loop. Two-phase cooling has demonstrated a huge potential that could tip balance in 

favor of a secondary loop. 

The remainder of this report will discuss potential advantages of these two-phase systems, briefly 

describe three competing two-phase cooling schemes, and describe one scheme in-depth (twophase microchannel coolers) that has the most potential to cool vehicular applications with 

current Army specifications in mind. In addition, we will discuss impediments to two-phase 

microchannel cooling techniques, examining steady-state research and introducing a relatively 

unmitigated obstacle, two-phase microchannel transient behavior. 

2. Two-Phase Cooling Background 

Two-phase cooling refers to intentionally using a cooling fluid in a manner where at least a 

portion of the fluid is transformed into vapor upon heating, thereby resulting in a gas/liquid 

mixture in a portion of the cooling loop. The boiling event generally occurs when the 

temperature of the heat acquisition surface exceeds the liquid’s saturation temperature, or boiling 

point. At this temperature, the vapor pressure of the liquid is equal to the pressure of the 
5 

environmental surroundings, thus allowing vapor bubbles to form at the solid-liquid interface, 

grow, and eventually detach from the surface. After this point, multi-phase flow exists in the 

system until the vapor is either separated from the liquid flow or is cooled to the point of 

condensing back into a liquid; schematically, the components typically used (the filter and 

vapor/liquid separator are optional) to achieve a two-phase loop are presented in figure 1. A 

liquid may be induced to boil over a range of temperatures, depending on the pressure of the 

environment, where increasing system pressure will increase the nominal boiling temperature of 

the fluid. This provides an additional benefit for a system using two-phase cooling, whereby the 

boiling and condensation temperatures of enclosed coolant loops may be tailored by controlling 

system pressurization. 

Figure 1. Primary components of a two-phase cooling loop. 

A two-phase system has several potential advantages over a standard single-phase liquid cooling 

system. First, a fluid’s latent heat of vaporization has a fundamental limit that can be two orders 

of magnitude larger than the specific sensible heat of single-phase liquid cooling (28). Thus, the 

boiling effect provides the possibility of increased heat absorption per unit volume of fluid and 

higher heat acquisition effectiveness (i.e., the amount of heat absorbed by a unit of flow relative 

to its maximum theoretical capacity). This latent heat benefit is coupled with the improved 

convection due to buoyancy-driven bubble formation, multi-phase turbulation, and mixing that 

takes place within the heat transfer region. As shown in figure 2, the resultant heat transfer 

coefficients from two-phase flow can be an order of magnitude greater than equivalent singlephase forced or passive (natural) convection (28). A larger heat transfer coefficient translates 

directly into a larger cooling capability for the system. A typical rule of thumb for two-phase 

liquid cooling systems is that the heat of vaporization and convective benefits of boiling can 

Condenser 

Filter 

Heat source / electronics

Pump

Vapor/liquid 

Fluid lines
6 

yield a two-to-four-times increase in heat removal capability (14). This performance projection 

is currently limited by safety factors and system derating associated with uncertainties of twophase cooling; by developing a better understanding of the process, engineers may be able to 

take full advantage of the vast convection and latent heat improvements. 

Figure 2. Heat transfer coefficients possible with natural 

convection, single-phase liquid forced convection and 

boiling for various fluids; from (28). 

In addition to improved cooling capacity, two-phase cooling also has the potential benefit of 

providing a relatively isothermal cooling surface (possibly less than 1 °C) (14, 29, 30). Whereas 

sensible heating of a single-phase coolant creates a downstream temperature rise, inducing a 

lateral temperature gradient along the cooled surface, two-phase coolers use latent heat 

absorption for energy transfer, which occurs at nearly isothermal conditions. Because excessive 

temperature and cooling non-uniformity over the electronic device area can induce thermal 

stresses and potential failure, isothermal cold plates can greatly increase the functional reliability 

of attached components (5). 

Weight, volume, and cost reduction are additional potential benefits of two-phase cooling 

techniques. Due to the increased effectiveness of two-phase cooling, lower flowrates can be 

used to deliver cooling performance similar to high flowrate single-phase systems. The potential 

value of lower system flowrates includes: smaller pumps with lower pumping power 

requirements; smaller onboard fluid volume; smaller fluid reservoirs; smaller tubing; and a 

reduction in related packaging weight and volume. 

During the past several decades of two-phase cooling research, three primary cooling schemes 

have gained the most attention: spray cooling, jet-impingement, and microchannel cooling (28). 

It is worth noting that each of these techniques can be used for single-phase or two-phase 

operation. During startup conditions, even a two-phase system has to move through a liquidonly single-phase regime. Proper system design would permit successful transient passage 

through this regime to the steady two-phase design point. The following sections will discuss the 
7 

advantages and disadvantages of each of these techniques, mainly focusing on two-phase cooling 

characteristics. 

2.1 Spray Cooling 

In figure 3, the primary characteristics of a spray cooling configuration can be seen. Fluid is 

introduced to an orifice or nozzle offset from the heated surface. With appropriate driving 

pressure, the fluid leaving the orifice breaks up into individual droplets that impinge upon the hot 

surface. The fluid will form a film upon the surface, the thickness of which is dependent on the 

relative fluid removal rate. With sufficient heating of the hot surface, there will be bubble 

formation in the liquid film, and the system will be operating as a two-phase cooler. 

Figure 3. Primary components of a spray cooler. 

Figure 4 shows qualitatively the relatively even cooling distribution that spray cooling can 

produce, primarily as a result of the uniformity of the spray pattern on the heated surface. Along 

with the other previously mentioned potential benefits of two-phase cooling, this can provide a 

uniform heat flux and temperature across the electronics. This fluid delivery method not only 

increases the spatial uniformity of heat removal but also delays liquid separation and eventual 

burn-out during vigorous boiling (28). As discussed in more detail in section 3.2, liquid 

separation refers to a vapor blanket present between the coolant fluid and heat source and is one 

condition leading to potential system overheating and failure. For the most part, spray coolers 

also provide low flow rate requirements due to the precise fluid-delivery and heat removal 

mechanism. 
8 

Figure 4. Qualitative comparison for heat transfer coefficients of spray 

coolers and jet impingement coolers; from (28). 

Despite these advantages, spray cooling is considered excessively complex and unfeasible for 

many applications (1, 31). Most spray cooling demonstrations have used standard commercial 

nozzles not necessarily designed or optimized for cooling applications. These nozzles generally 

require very large driving pressures (10–45 psi or even higher) to induce droplet breakup (24), 

and even supposedly identical nozzles can exhibit a definite lack of spray pattern repeatability 

and heat transfer performance (18). In addition, erosion, corrosion, and nozzle clogging are 

primary concerns for practical implementation that may require design concessions that degrade 

thermal performance (18). 

The modeling and prediction of spray cooling performance also suffers from the sheer 

complexity of the fluidic and thermal behavior or the system. This complexity is due to the large 

number of parameters associated with spray cooling performance, many of which are hard to 

control or characterize in an experimental setup. In addition to parameters important for twophase cooling, in general, such as fluid and heated surface properties, other spray parameters 

include nozzle type and shape, droplet size, volume flux, spray angle, nozzle-to-surface distance, 

inclination with respect to gravity, and spray overlap—all of which can significantly affect 

cooling performance (32–46). 

Even though two-phase spray coolers have demonstrated drawbacks, research efforts have shown 

the potential to provide high-heat-flux electronics cooling. In a study by Lin and Ponnappan (24) 

using FC-87 and methanol, heat fluxes of 91.5 W/cm2

 and 490 W/cm2, respectively, were 

demonstrated before overheating and physical burn-out of the electronics occurred. Mudawar et 

al. (7) developed a spray cooler as part of the U.S. Department of Energy’s (DOE’s) Power 

Electronics and Electric Machines program for developing hybrid-electric vehicles. A two-phase 

spray cooling module was implemented using HFE-7100 dielectric coolant fluid to safely 

dissipate 150-200W/cm2

 while maintaining the chip temperature below 125 °C. Similarly, Turek 
9 

et al. (47) developed a two-phase spray cooling module for a 180 kW Semikron Insulated Gate 

Bipolar Transistor (IGBT) module using 100 °C 50/50 PGW mixture at a pressure drop of 40 psi, 

providing a heat flux of 165 W/cm2

. In a laboratory setting, Yang et al. (48), with a multi-nozzle 

spray cooler using water as the working fluid, reached a heat flux of ~1000W/cm2

. 

2.2 Jet-impingement Cooling 

Jet impingement can be implemented in three basic forms: free jet (liquid jet in a vapor or 

gaseous ambient), submerged jet (liquid jet in a liquid ambient), and confined jet (liquid jet 

confined between the nozzle and heat source). From figure 5, the primary components of a free 

jet jet-impingement cooler can be seen. With sufficient driving pressure, a continuous jet of 

fluid is forced through a jet nozzle that impinges upon the heated surface. With adequate 

heating, bubbles will form and the jet-impingement cooler will operate as a two-phase cooler. 

Figure 5. Primary components of a jet-impingement cooler. 

The fluid flow will take two forms—an impingement zone, where the fluid is colliding with the 

hot surface, and a flow region, where the liquid is flowing from the impingement zone to 

adjacent surfaces. The multiple fluid regions (impingement and flow) give rise to the jet profile 

shown in figure 4, with the impingement zone having a much larger heat transfer coefficient than 

the flow region. Furthermore, figure 4 demonstrates jet-impingement’s potential to provide 

significantly improved localized heat transfer coefficients compared to spray cooling. Since jets 

concentrate most of the cooling potential in the impingement zone, multiple jets are commonly 

used to create several impingement zones to cool larger areas (28, 48, 50). Because the fluid is 

continuous, unlike droplets in spray cooling, the fluid dynamics are a bit simpler. Modeling 
10 

using current Computational Fluid Dynamics (CFD) approaches is possible, but still nontrivial, 

and requires user-defined functions (51, 52). 

However, the concentration of the heat transfer coefficient and associated heat removal within 

the impingement zone can cause significant temperature gradients within the device. This 

temperature non-uniformity can cause thermal stress in the electronics and potential failure. Jetimpingement is also prone to separation of the wall jet with vigorous boiling, leading to dry-out 

and eventual physical burn-out of the electronics (53). Furthermore, jet-impingement is 

considered an aggressive form of cooling due to the large impact momentum imparted on the 

heat source; the impact force may be enough to damage delicate devices requiring additional 

system implementation considerations (28).

Two-phase jet-impingement system designs, despite known shortcomings, have also 

demonstrated the ability to cool high-heat-flux electronic modules. A preliminary test by 

Copeland (49) demonstrated heat fluxes as high as 369 W/cm2

 in a two-phase pin fin array, using 

multiple slot nozzle impingement with FC-72 fluorocarbon liquid and moderate flowrates 

ranging from 0.075 to 0.6 l/min. Phase-change jet-impingement cooling was also demonstrated 

by Mudawar (28) on the SEM-E BTPFL-C3 avionics Clamshell Module. By using direct twophase jet-impingement and FC-72 dielectric fluid, the BTPFL-C3 performance exceeded 10 kW, 

and heat fluxes as high as 255 W/cm2

 were demonstrated. Bhunia et al. (25) demonstrated power 

densities up to ~1000 W/cm2

 in a laboratory setting tailored towards high power density 

electronic cooling. 

2.3 Microchannel Cooling 

Shown in figure 6 are the primary components of a microchannel cooler. Electronic devices are 

typically attached to the topside of the microchannel cooler, with the backside left open for 

microchannel fabrication. With appropriate driving pressure, the fluid is forced through large 

surface area microchannels, and heat is transferred from the topside electronics, through the 

microchannels, and into the fluid. Similar to spray-cooling and jet-impingement, microchannel 

systems were initially designed as single-phase systems, but with sufficient heating the system 

will operate as a two-phase cooler. 
11 

Figure 6. Primary components of a microchannel cooler. 

Two-phase microchannel systems have attracted much attention as a method for high-heat flux 

electronic cooling. While it has been shown that flow through large surface area microchannels 

boosts the heat transfer coefficient in single-phase flow (16, 54), additional gains in the heat 

transfer coefficient can be obtained by allowing the liquid to vaporize along the walls of the 

microchannel. These systems also possess attributes like compactness and minimal coolant usage 

that make them attractive for electronics cooling applications. 

However, two-phase microchannel heat sinks have demonstrated shortcomings that will 

necessitate extensive engineering knowledge to overcome in order to develop effective cooling 

systems. Two-phase microchannel heat sinks have significant steady-state and transient flow 

and pressure uncertainties, making them prone to dry-out and critical heat flux overshoot that can 

cause overheating and electronic burn-out. Boiling incipience and conjugate effects, such as 

nonuniform fin temperature, are also contributing factors to the uncertainty of two-phase 

microchannel cooling, limiting the practical application of such systems (55, 56). 

Despite these impediments, two-phase microchannel designs have demonstrated the potential to 

provide the cooling necessary for high heat flux electronic systems. One example is a study 

performed by Lee and Mudawar (13) with R134A and HFE1700 direct and indirect 

microchannel cooling systems; the study reports heat fluxes as high as 840 W/cm2

 without 

overheating or electronic burn-out. As an extreme, a study by Mudawar et al. at Purdue 

University (56, 57) have confirmed ultra-high heat flux values as large as 27,600 W/cm2

 with 

subcooled water flow at high velocity in small diameter micro-tubes (D<1mm). While this 

performance may not be feasible with Army vehicular flowrate and pressure requirements, this 

study demonstrates the capability of two-phase cooling and reiterates the potential for such 

cooling schemes. 
12 

2.4 Technology Comparison 

The three previously described two-phase cooling technologies have demonstrated the capability 

of providing high levels of cooling using elevated temperature coolants. The successful use of 

any of these techniques would potentially enable high heat flux electronics to operate on an 

elevated temperature coolant loop. Common to all technologies are operational uncertainty, the 

risk of instability under steady or transient loads, and the presence of a sudden, critical failure 

point. This poor understanding leads to increased margins of safety, forcing operation of the 

cooling component far from its peak performance, and reducing the net benefit of two-phase 

cooling schemes from the start. 

Specific to each technique, however, are particular drawbacks that might make Army vehicle 

implantation problematic. With respect to spray cooling, the high driving pressure is the primary 

stumbling block to system compatibility. While two-phase cooling techniques all have tended to 

exhibit higher pressure drops than single-phase equivalents, a large driving pressure is the 

fundamental mechanism to spray formation. Because jets and channels are continuum flows, 

they generally can be operated over a wide range of flowrates, and the system compatibility 

challenge becomes improving performance at lower pressures. A spray component would require 

system level redesign to support the pressure requirement necessary for adoption into the 

previously mentioned cooling loop architecture. 

Both the jets and the microchannels seem capable of operating over a similar pressure range. The 

jets have the potential for high convection coefficients, but this appears to be highly localized to 

the impingement region of the jet. While jet arrays can reduce this effect to some degree, an 

array configuration introduces additional complexity in terms of fluid routing and removal. 

Proper manifolding of such a jet configuration to manage fluid delivery and removal could be a 

nontrivial design and assembly problem (58). While there have been numerous single-phase 

microchannel studies examining complex manifolding techniques (59), their general purpose is 

to reduce the temperature rise effect of sensible heating, which two-phase approaches already 

mitigate via latent heat effects. Thus, standard microchannel designs with relatively simple fluid 

paths could provide appropriate cooling performance within the confines of our system limits. 

It is also worth mentioning that the microchannel configuration can be somewhat simpler from 

an academic sense, with fewer geometric variables and little interaction between individual 

elements, making it a more tractable problem. With these factors in mind, two-phase 

microchannel coolers represent the most promising option for high heat flux power electronics 

cooling goals in Army vehicle applications. 
13 

3. Two-phase Microchannel Flow Obstacles 

In general, two-phase flows are complex physical processes dependent on a variety of variables: 

channel shape, flow direction, fluid compressibility, interfacial tension forces, fluid wetting 

characteristics, momentum, pressure, temperature, fluid velocity, and channel orientation. These 

can all contribute toward various liquid-vapor two-phase flow patterns and regimes (60). As 

demonstrated by extensive literature (61–78), the mechanisms controlling two-phase pressure 

drops, heat transfer coefficients, and critical heat flux are inherently related to flow regimes, thus 

making flow regime prediction a key component of research and two-phase heat transfer design. 

In addition to steady-state flow concerns, the current understatement of two-phase transient 

behavior is also presenting a hindrance, limiting practical system implementation. If dynamic 

startup and transient conditions are left unmitigated in such systems, not only will system 

stability decrease, but the potential for system failure will increase (79, 80). The following 

sections are aimed towards describing the nature of these problems. 

3.1 Microchannel Two-phase Steady-state Flow Regimes 

An extensive knowledge base surrounding steady-state two-phase flow exists and, while it is 

hard to predict the stability with specific design parameters in mind, it typically follows the 

progression outlined in the following section. Figures 7a and 7b show the flow regimes for twophase microchannel flow in horizontal and vertical orientations, respectively. With a subcooled 

inlet, where the temperature of the entering fluid is lower than the saturation temperature, singlephase forced liquid flow continues for a finite length until sensible heat absorption raises the 

temperature enough for subcooled boiling to commence along the wall. The point between 

liquid-forced convection and subcooled boiling is known as incipent boiling and indicates the 

condition where boiling can begin. There are two regions in subcooled boiling—the first where 

the core liquid is very subcooled and boiling only occurs on the walls, and the second where the 

fluid has a higher temperature, but still below the saturation temperature, and vapor bubbles can 

be found in the core of the microchannel. This difference occurs because the bubbles 

immediately condense in the liquid core of very subcooled liquid; a liquid core closer to the 

saturation temperature, however, allows bubbles to persist in the bulk flow without condensing. 

The interface between these two subcooled boiling regimes is the point of net vapor generation. 

The next bulk region is the saturated boiling regime, which consists of bubbly flow, slug (plug) 

flow, and annular flow. In the saturated boiling region, the vapor quality can range from 0 (0% 

vapor and 100% liquid by mass) to approximately 1 (100% vapor and 0% liquid). 

3.2 Thermal Failure and Critical Heat Flux (CHF) 

From a design perspective, there are thermal implications associated with the flow regimes 

discussed above. As explained by Incropera et al. (21), and indicated in figure 7a, the transition 
14 

from subcooled boiling, to bubbly flow, slug flow, and annular flow represents increases in 

vapor quality, heat transfer coefficient, and thermal performance. The annular flow region 

persists until wall dry-out and, eventually, vapor forced convection occur. This transition, and 

subsequent dry-out, results in severely reduced heat transfer coefficients due to the much lower 

convective performance of vapor compared to liquid. Because the system will be able to 

dissipate less heat after a reduced convection coefficient due to dry-out, temperatures in those 

regions can be much higher and result in system overheating and burn-out. 

(a)

(b) 

Figure 7. Ideal microchannel 2-phase boiling flow regimes in (a) horizontal and (b) vertical 

orientations; from (60). 
15 

As discussed previously and shown in figure 8, there is a point within the microchannel system 

where dry-out and possibly critical thermal failure can occur. This condition will be dependent 

on the amount of heat being added to the system, and the thermal limit inducing this condition 

has been described as the critical heat flux (CHF) or boiling crisis. Referring to figure 8, Point A 

indicates the transition from single-phase to boiling heat transfer. As expected with the latent 

absorption and improved convection due to boiling, the slope of the curve increases greatly to the 

right of Point A and higher heat flux is obtainable with minimal wall temperature rise. Point C 

indicates the location where excess heating begins to causes partial or total dry-out, and a sudden 

decrease in heat transfer coefficient is observed. This location typically occurs in the saturated 

annular flow regime, but it can also be seen in subcooled boiling, where excess heat causes a 

vapor blanket to form on the walls of the channel even though the actual vapor quality for the 

entire channel is still relatively low. While operating the system close to CHF typically yields 

annular flow and higher heat transfer coefficients, exceeding the CHF limit would cause 

overheating of the channels and potential electronic failure. In general, the design objective in 

creating a two-phase microchannel cooler is to induce boiling at the earliest point in the 

microchannel without encountering dry-out at the end of the channel—or, in other words, 

without exceeding the critical heat flux for that flow. However, the inability to predict flow 

regimes and subsequent CHF design values has required a factor of safety be used in the design 

of any of these cooling systems, and operating too close to any perceived non-recoverable failure 

point is unacceptable. As shown on figure 8, a factor of safety would move system operation 

down on the curve and provide less than optimized heat flux. The power derating that would 

currently be needed to guarantee the system stay below these estimated critical failure points 

could likely render the cooling solution less effective than current single-phase options. Research 

must be directed at higher fidelity prediction of these failure limits and better design for stable 

operation. 
16 

Figure 8. Critical Heat Flux for Water @ 1atm. Point A is the onset of boiling and graph 

follows typical flow progression up to annular flow and CHF at Point C; from (21). 

It is worth noting that since microchannels are typically cut into materials with moderate to large 

thermal conductivities, local critical heat flux may not always cause failure of the entire cooling 

system. From a practical perspective, vapor formation and dry-out of a portion of the 

microchannel cooler will cause local overheating. Due to a large thermal conductivity, however, 

the heat can be redistributed to other other areas of the microchannel heat sink that can 

accommodate the heat dissipation. Therefore, the reliability and robustness of a two-phase 

microchannel cold plate can be increased by using materials with high thermal conductivities. 

In microchannel applications where the fluid is entering in a liquid state, it is important to 

determine if there are any barriers to vapor formation; vapor formation indicates the onset of 

boiling and is necessary in a two-phase cooling system to achieve the thermal benefits associated 

with boiling. It has been presented in the literature (28, 81, 82) that vapor formation and boiling 

inception is a key factor to system performance and critical heat flux increases. By creating 

surface imperfections or nucleation sites, boiling enhancement in the form of micro-porous 

surfaces, painted diamond particles, surface roughening, and painted silver flakes have 

demonstrated boiling inception temperatures 75–85% lower than smooth surfaces with system 

CHFs as much as 224% higher (82–84); this would indicate a shift up and left on the curve in 

figure 8. Also in figure 8, improvements in CHF represent lower wall superheat necessary for 

boiling inception and higher allowable heat flux before potential failure occurs. As an avenue 

for cooler performance enhancement, engineers have explored combining microchannels with 

surface modification techniques previously discussed for further improvement in boiling 

CHF Improvements 

Safety factor reduced thermal performance 
17 

inception and critical heat flux. It is important to note that while smaller microchannels and 

rougher surfaces improve the thermal transport qualities, they also increase the pressure drop 

across the device (85, 86). For Army applications, significant design attention would be 

necessary to balance CHF improvements and associated pressure drops. 

3.3 Channel Orientation 

There has been little experimental attention paid to inclined and declined microchannel flows, 

despite obvious technical importance; many Army vehicles can be expected to climb significant 

grades, therefore, the associated electronics and coolers would encounter these inclines as well. 

As described by Hewitt (63), inclined and declined two-phase microchannel flows can be 

expected to demonstrate qualities of horizontal and vertical orientations combined. 

Due to buoyancy effects, gravity can shift the distribution of the different fluid phases in the 

channel. Generally, horizontal microchannel flows demonstrate asymmetry due to gravity, 

causing stratification with the liquid at the bottom of the channel and vapor to occupy the top; 

this is illustrated in figure 9a, which is in contrast to the more uniform liquid-vapor distribution 

shown for a vertical configuration in figure 9b. This asymmetry and stratification causes thermal 

complications—possible intermittent drying and rewetting of the upper surface of the tube over 

short distances, or total dry-out of the upper surface in long channels that can cause overheating 

and burn-out of the channel and electronics (87–91). This scenario is exacerbated by potential 

non-uniform conductive heat flow around the perimeter of a microchannel, causing variation in 

the boiling regime and associated heat transfer coefficient within the channel (55, 92, 93). To 

mitigate these effects, studies have shown that higher inlet liquid velocities and high thermal 

conductivity microchannel materials lead to more symmetric flow and reduced non-uniform heat 

flow (60). 

 (a) (b) 

Figure 9. Non-Ideal microchannel 2-phase boiling flow regimes in (a) horizontal and (b) vertical orientations; 

from (60). 

3.4 Microchannel Two-phase Transient Effects 

As discussed previously, significant attention has been paid to steady state two-phase flow 

regimes. However, startup and dynamic conditions in a two-phase cooling system are very 

complex transient phenomena that need to be considered. Specifically, special attention must be 
18 

paid to variations in heat load, mass flowrate, electronic component temperature, and coolant 

fluid temperature. Two studies exploring the transient and startup of a novel pumped two-phase 

cooling loop were presented (79, 80), demonstrating acceptable performance in the form of 

stable temperature, pressure, and flow. While there has been some experimentation concerning 

two-phase transients, there is an overall lack of documented experimental results for two-phase 

microchannel transient performance. The result is systems designed with larger safety margins, 

reducing the net benefit of two-phase cooling schemes. 

3.5 Microchannel Flow Summary 

In general, several statements can be made about the status and future path of two-phase 

microchannel cooling techniques. While there are a significant number of studies in 

microchannels, there is no proven consistency for the same fluids under the same test conditions. 

This may be in large part due to the inconsistent and subjective visual-only methods used to 

characterize flow regimes. While several review papers outlining prediction of two-phase 

microchannel heat transfer characteristics and flow maps are available (60, 94), widely accepted 

theoretically based flow regime flow maps have not been established (60, 95). Research must be 

directed in several directions to provide systematic knowledge and promote higher fidelity 

prediction of two-phase flow in microchannels: 

1. Effects of channel shapes, channel orientation, fluid properties, and other physical 

variables. 

2. Steady-state flow regimes and associated failure limits. 

3. Transient conditions to better design for stable and dynamic operation. 

4. Two-phase Macro/Microchannel Commercial Success 

Two-phase macrochannel coolers are being implemented as an intermediary step between 

conventional single-phase methods and future two-phase microchannel designs. At the cost of 

reduced thermal performance, macrochannel designs have a more extensive research basis and 

are currently considered less complex and easier to implement than their higher-performing 

microchannel counterparts. With the design philosophy of “simplicity and effectiveness” in 

mind, Parker Hannifin has developed a high-voltage power system for military vehicle 

applications using a two-phase macrochannel cooling technology (14). They are currently past 

the research and development stages with their technology and are deep into commercialization 

and shipping. 

Figure 10 is a photograph of a Parker Hannifin 450A IGBT two-phase cooling module. Their 

system uses R134A dielectric fluid as the coolant, a low flowrate positive displacement pump 

(not shown), specially designed two-phase corrugated copper cold plates, and a vapor condenser 
19 

(not shown). R134A was chosen because it is a well-characterized refrigerant with widely 

commercially available pumps and components; as previous fluid selection information implies, 

other fluids may be used with appropriate pump and component modifications. 

Figure 10. Two-phase Parker Hannifin 450A IGBT cooling module; from (14). 

The performance of this module is characterized by a power dissipation comparison to other 

cooling technologies. As outlined in table 1, five different cooling schemes were employed to 

observe the relative performance of the two-phase cooler. 
20 

Table 1. Parker-Hannifin inverter test cases; from (14). 

Figure 11 shows that the two-phase cooling technique (VDF Copper, vaporizable dielectric fluid) 

is superior in cooling capacity to all other tested single-phase cooling methods. Not only does 

the VDF Copper two-phase cooling method have a significantly reduced flowrate (compare 0.4 

to 2 GPM in table 1), the cooling capacity is about 40% better than the next highest performing 

single-phase cooling technique (14). 
21 

Figure 11. Cooling Performance of the two-phase Parker Hannifin cooler compared to 

single-phase cooling techniques; from (14). 

Pumping power consumption, an advantage mentioned earlier, is a major benefit when 

considering a two-phase cooling system. The pumping power necessary to cool a 1 kW heat 

load with the R134A two-phase copper heat sink was measured to be 12 W. With an equivalent 

load and similar temperature limits, a single-phase water aluminum pressure-fitted cold plate 

required 295 W of pumping power to match the cooling potential of the two-phase method. This 

system also has the added benefit of an isothermal heat sink temperature, which reduces the 

thermal stress on silicon chips and improves reliability (14). 

Compared to standard single-phase cooling techniques, the above commercial example 

demonstrated the potential for isothermal cold plates, pumping power reduction, flowrate 

reduction, and improved heat dissipation. Despite these achievements, areas of improvement 

relating to CHF need to be explored. As previously indicated, approaching CHF in annular flow 

with high vapor qualities would be beneficial to increase heat flux, reduce flowrate, and 

minimize size and weight of associated components. However, this module was designed 

around selecting an appropriate vapor quality maximum value (70%) as not to exceed CHF and 

prevent dry-out from occurring within the cold plate (14). Aforementioned steady-state 

uncertainties and transient conditions played a large role in this module’s cautious design, and 

improved research attention in these areas should be a major focal point to further increase 

predictability and performance of two-phase systems. 
22 

5. Conclusions and Recommendations 

The difficulties in designing a hybrid electric propulsion system for Army vehicular applications 

have consistently called for advanced thermal management technologies. However, single-phase 

forced-convection liquid cooling has reached a mature state in the academic and commercial 

worlds, and it is highly unlikely that the order-of-magnitude heat flux improvements desired for 

future systems will be obtainable by this technology under the current set of platform constraints. 

The key limitations can be summarized as: 

1. Minimum allowable orifice size—particles and sediment in the primary engine loop limit 

cooling scheme geometries and may cause clogging. 

2. Fluid limitations—the primary cooling loop has been designed for an EGW coolant, which 

is not an ideal fluid for use in electronics cooling. 

3. Pumping inflexibility—fluid delivery requirements can not be optimized for the electronic 

cooling components due to engine coolant flow and pressure specifications. 

4. Single-phase fluid limitation—because the system is not designed to accommodate mixed 

gas/liquid flow, there is currently little opportunity to implement two-phase flow. 

This paper has examined electronics cooling on the Army vehicle platform, with particular focus 

on the factors limiting the use of two-phase microchannel cooling technology. Based on the 

platform specific limitations just described and the current unknowns in the academic 

community on two-phase microchannel cooling, the following research directions are 

recommended for achieving the thermal performance necessary for future systems: 

1. Determine effects of EGW additives—additives may significantly influence two-phase 

cooling performance and behavior compared to “pure” EGW. 

2. Steady-state two-phase flow risk reduction—reliable and accurate models of flow stability 

and CHF must be established to enable adoption in real systems without significant 

performance-reducing safety margins. 

3. Transient two-phase flow risk reduction—the effects of two-phase cooling under dynamic 

conditions must be understood to improve system reliability. 

4. Study vehicle architectures compatible with two-phase cooling—examining the systemlevel costs and benefits of multi-loop architectures would help identify potential Army 

vehicle cooling solutions. 

Single-phase cooling techniques do not possess the cooling capability necessary to enable future 

Army power-dense hybrid-electric vehicle applications, especially with the requirements for 
23 

elevated coolant loop temperatures. Despite demonstrated performance, two-phase cooling has 

significant technical uncertainties, causing a level of implementation risk. Only by reducing the 

risk of two-phase cooling solutions while maintaining the promised power density improvements 

will system builders be willing to consider the technology for implementation in future 

platforms. 
24 

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33 

Appendix A. Typical Coolant Fluids and Associated Properties 
34 

List of Symbols, Abbreviations, and Acronyms 

ARL Army Research Laboratory 

CFD Computational Fluid Dynamics 

CHF Critical Heat Flux 

DoD Department of Defense 

DOE Department of Energy 

EGW Ethylene Glycol and Water 

IGBT Insulated Gate Bipolar Transistor 

NASA National Aeronautics and Space administration 

PGW Propylene Glycol and Water 

Sodium MBT Sodium Mercaptobenzothiazole 

SWAP Size, Weight, and Power 

TARDEC Tank Automotive Research, Design, and Engineering Center 

VDF Vaporizable Dielectric Fluid 
35 

NO. OF 

COPIES ORGANIZATION 

 1 ADMNSTR 

 ELEC DEFNS TECHL INFO CTR 

 ATTN DTIC OCP 

 8725 JOHN J KINGMAN RD STE 0944 

 FT BELVOIR VA 22060-6218 

 1 CD OFC OF THE SECY OF DEFNS 

 ATTN ODDRE (R&AT) 

 THE PENTAGON 

 WASHINGTON DC 20301-3080 

 1 US ARMY RSRCH DEV AND ENGRG 

 CMND 

 ARMAMENT RSRCH DEV & ENGRG 

 CTR 

 ARMAMENT ENGRG & TECHNLGY 

 CTR 

 ATTN AMSRD AAR AEF T 

 J MATTS 

 BLDG 305 

 ABERDEEN PROVING GROUND MD 

 21005-5001 

 1 PM TIMS, PROFILER (MMS-P) 

 AN/TMQ-52 

 ATTN B GRIFFIES 

 BUILDING 563 

 FT MONMOUTH NJ 07703 

 2 US ARMY TARDEC 

 RESEARCH BUSINESS GROUP 

 GVPM 

 POWER PLANT INTEGRATION 

 TEAM 

 ATTN M GORYCA 

 ATTN O TARNAVSKY 

 6501 EAST 11 MILE RD 

 WARREN MI 48397-5000 

 1 US ARMY INFO SYS ENGRG CMND 

 ATTN AMSEL IE TD A RIVERA 

 FT HUACHUCA AZ 85613-5300 

 1 COMMANDER 

 US ARMY RDECOM 

 ATTN AMSRD AMR 

 W C MCCORKLE 

 5400 FOWLER RD 

 REDSTONE ARSENAL AL 35898-5000 

NO. OF 

COPIES ORGANIZATION 

 1 OFFICE OF NAVAL RSRCH 

 ATTN CODE 331 M SPECTOR 

 875 NORTH RANDOLPH ST 

 ARLINGTON VA 22203-1995 

 1 US GOVERNMENT PRINT OFF 

 DEPOSITORY RECEIVING SECTION 

 ATTN MAIL STOP IDAD J TATE 

 732 NORTH CAPITOL ST NW 

 WASHINGTON DC 20402 

 1 UNIV OF MARYLAND 

 ATTN A BAR-COHEN 

 2181B GLENN L. MARTIN HALL 

 BLDG 088 

 COLLEGE PARK MD 20742 

 1 US ARMY RSRCH LAB 

 ATTN RDRL CIM G T LANDFRIED 

 BLDG 4600 

 ABERDEEN PROVING GROUND MD 

 21005-5066 

 6 US ARMY RSRCH LAB 

 ATTN IMNE ALC HRR 

 MAIL & RECORDS MGMT 

 ATTN RDRL CIM L TECHL LIB 

 ATTN RDRL CIM P TECHL PUB 

 ATTN RDRL SED E B MORGAN 

 ATTN RDRL SED E D SHARAR 

 ATTN RDRL SED E N JANKOWSKI 

 ADELPHI MD 20783-1197 

TOTAL: 18 (1 ELEC, 1 CD, 16 HCS) 
36 

INTENTIONALLY LEFT BLANK.