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Applied Thermal Engineering 60 (2013) 316e324

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Applied Thermal Engineering

journal homepage: www.elsevier .com/locate/apthermeng

Demonstration of a wearable cooling system for elevated ambienttemperature duty personnel

Timothy C. Ernst a, Srinivas Garimella b,*

aAdvanced Engineering, Cummins, Inc., Columbus, IN 47201, USAb Sustainable Thermal Systems Laboratory, George W. Woodruff School of Mechanical Engineering, Georgia Institute of Technology, Atlanta, GA 30332, USA

h i g h l i g h t s

� Wearable cooling system for use in elevated temperature environments developed.� 2.0 L of fuel powers engine that runs compressor to provide 5.7 h of cooling.� The 0.318 � 0.273 � 0.152 m cooling system has a mass of 5.31 kg.� Cooling of up to 300 W at ambient temperatures of 37.7e47.5 �C demonstrated.

a r t i c l e i n f o

Article history:Received 28 July 2011Accepted 11 June 2013Available online 25 June 2013

Keywords:Heat pumpReciprocating compressorR134aMiniaturization

* Corresponding author. Tel.: þ1 404 894 7479.E-mail address: sgarimella@gatech.edu (S. Garime

1359-4311/$ e see front matter � 2013 Elsevier Ltd.http://dx.doi.org/10.1016/j.applthermaleng.2013.06.01

a b s t r a c t

Awearable cooling systemwas developed for use in elevated temperature environments by military, fire-fighting, chemical-response, and other hazardous duty personnel. The cooling system consists of anengine-driven R134a vapor compression system assembled in a backpack configuration, coupled with acooling garment containing refrigerant lines. A 2.0 L fuel tank powers a small-scale engine that runs acompressor fabricated in house. The overall cooling system, including the wearable evaporator, had amass of 5.31 kg and measured 0.318 � 0.273 � 0.152 m. Controlled environment tests determined systemperformance over a range of ambient temperatures (37.7e47.5 �C), evaporator refrigerant temperatures(22.2e26.1 �C), and engine speeds (10,500e13,300 RPM). Heat removal rates of up to 300 W, which is thecooling rate for maintaining comfort at an activity level comparable to calisthenics or moderate exercise,were demonstrated at an ambient temperature of 43.3 �C. The system consumed 1750 W at a fuel flowrate of 0.316 kg h�1 to provide a 178 W of cooling for 5.7 h.

� 2013 Elsevier Ltd. All rights reserved.

1. Introduction

Reducing thermal stresses for hazardous-duty personnel canincrease productivity and allowable durations of missions, reducefatigue, and lead to a safer working environment. Thermal stressescan be reduced by means of portable cooling systems in elevatedtemperature environments. Potential users of such systems includethe military, fire-fighters, and other hazardous duty and chemical-spill response personnel. Semi-portable cooling systems that con-nect to centralized cooling systems are available; however, thisrequires the user to be tethered to a fixed point at a certain radius,thus restricting the range of motion. Therefore, a self-contained,portable cooling system was developed in this study. The threemajor challenges in the development of such portable cooling

lla).

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systems are: providing input power, miniaturization of compo-nents, and heat dissipation to ambient air. For portable operation, asystem that is as lightweight and compact as possible is essential.The system is intended to be used independently from any external(stationary) power or input, which requires that the cooling systembe capable of providing input power, as well as operating inde-pendently in the environment in which the user functions. Thisenvironment is at an elevated temperature, which further increasesthe challenge of heat removal and rejection.

2. Previous work

A variety of methods have been investigated to provide portablecooling of personnel working in high thermal stress conditions. Adetailed survey of such techniques was provided recently by theseauthors [6,7]. A brief overview is provided here. Although ice packsor other phase-changematerials are simple and reliable methods ofcooling, for extended operation, transportation of replacement

T.C. Ernst, S. Garimella / Applied Thermal Engineering 60 (2013) 316e324 317

packs is cumbersome [5]. A latent heat of fusion of 335 kJ kg�1

implies that 6.4 kg of ice are required to provide 300 W of coolingfor only 2 h. Adsorption systems lower vapor pressure and evapo-rate water at the desired temperature, resulting in a cooling effect.But they require components for the storage of both desiccant andwater, which for extended cooling periods, yield a large and heavysystem. An adsorption system using calcium oxide as the desiccantweighed 19.9 kg and had dimensions of 0.602 � 0.188 � 0.335 m[8]. Such systems have lag times between startup and the coolingeffect of 30e45 min [8]. Absorption cycles require heat input, andthe shaft work input to the pump ismuch less than that required forvapor compression. The heat input can be obtained using com-bustion of a liquid fuel, which has energy densities much higherthan that of batteries. But these cycles require higher complexitythan other systems, and also require large heat rejection rates to thesurrounding atmosphere.

When considering the design of cooling garments, the temper-ature of the garment must be low enough to enable heat transferfrom the body, but not so low as to feel uncomfortable. For in-dividuals at rest, the average comfortable skin temperature is 33 �C[11]. Cooling using air flow inside a garment removes heat byconvection and also through the evaporation of perspiration at lowhumidities. But using air cooling in water- or refrigerant-basedsystems would require an additional air-coupled heat exchangerand an additional temperature penalty to cool the air. It is alsodifficult to work in pressurized air-cooled garments [11], requiringthe user to work harder, thus increasing the cooling duty. Ref. [5]defined efficiency and the effectiveness as performance measuresfor cooling suits using air, water and ice. The efficiency is theamount of cooling per unit area, or heat flux, and the effectivenessis the total cooling of a particular area of the body. Their findingsshowed that cooling of the torso was more effective than otherparts of the body, due primarily to the larger surface area. The headwas shown to be the most efficient area; however, due to thelimited area, it was less effective. Ref. [14] recommended percent-ages of tubing to be allotted for maximum effectiveness in a fullbody cooling suit (torso 30%, head and neck 25%, thighs 17%, upperarms 15%, calves 8%, and forearms 5%). Ref. [10] tested the body’sresponse using a water-cooled three-layer cotton vest covering 20percent of body area lined with 2 mm diameter latex tubes cooledby an insulated ice-pack. The vest maintained comfortable bodycore and skin temperatures, but they also found that the suit itselfcan impede the body’s natural temperature regulation through theevaporation of perspiration. They recommended that the garmentshould wick away the perspiration. Ref. [12] tested flexible micro-channel heat exchangers made from a 0.2 mm thick heat-sealablepolyimide film, which are well suited to being used in a coolinggarment. They also withstand high pressures, which means thatrefrigerant can be routed to the garment without an additionalwater loop.

The above discussion shows that different methods have beeninvestigated for portable cooling. Some studies [13] have over-estimated fuel consumption and weight of engine-driven vaporcompression systems (11.3 kg, with fuel consumption> 0.454 kgh�1). Suchpredictions are not based on results fromactualexperiments on such systems. The present work therefore un-dertakes the design, fabrication, testing of a compact, wearableengine-driven vapor-compression system for portable coolingapplications.

3. System design considerations

For an acceptably long mission duration, input power must besupplied for a commensurate time, increasing the total energysupply carried by the user. Refs. [6,7] conducted a detailed analysis

of the available choices for power sources for portable coolingsystems for a nominal cooling load of 400 W. Based on an energydensity of 137.5 Wh kg�1 [2] for Lithium-Ion batteries, they foundthat batteries would be too massive and bulky for missions >1 h.Similarly, a fuel cell was found to have an initial weight of 6 kg [4],which is too large to compensate for its lower fuel requirement of0.054 kg h�1. This can be compared to a small-scale engine esti-mated to weigh w1 kg, with a fuel mass flow rate of 0.314 kg h�1

[1], even for durations well past 10 h. Despite the noise and exhaustproduced by engines, and their lower reliability, an engine waschosen as the power source for this study due to the significantadvantages in energy density, which enables long missions.

A vapor compression cycle was chosen for the cooling systembased on the fewer components required and the lower complexitycompared to absorption systems. R134a was chosen as the workingfluid because of the relatively high cycle efficiencies among thefluid alternatives considered [6,7], its ready availability and lowcost. The vapor compression cycle necessitated the development ofa miniaturized compressor because of the limited availability ofsmall-scale refrigerant compressors.

An air-coupled heat exchanger mounted on a backpack, and awearable condenser, were considered for heat rejection to theambient [6,7]. The exterior of the garment, separated from the in-ner evaporator by insulation, was analyzed as a condenser thatwould reject heat to the surrounding air passively. Although an air-coupled condenser would necessitate the use of a fan and addi-tional input power, the increased cooling load due to the parasiticheat gain into the evaporator from the wearable condenser wasfound to requiremuchmore (58%) input power. Thus, a systemwitha fan-cooled condenser built into the backpack was chosen. Acondenser air flow rate of 0.12 kg s�1 was chosen based on thetradeoff between fan power and compressor power. (Higher airflow rates increase fan power, but decrease condenser temperatureand pressure, leading to lower compressor work.)

The system chosen here was first modeled using thermody-namic cycle and heat exchange calculations, which determinedsystem size and energy input requirements. From this analysis, theappropriately sized air compressor was selected and modified in-house into a refrigerant compressor. Validation tests were firstconducted on a stand-alone compressor test stand [7]. Subse-quently, the overall systemwas assembled and tested to determinesystem performance at various combinations of ambient temper-ature, evaporator temperature and engine speed. Finally, the sys-tem developed in this manner was assembled into a compact,wearable cooling system, as described below.

4. Compressor development

A small-scale off-the-shelf (Target Corp. HC-5320284 12-voltDC) portable air compressor was used as a starting point andmodified considerably to function as a refrigerant compressor forthe portable cooling system. The compressor (Fig. 1) had a pistondiameter of 19 mm and a stroke of 17 mm, yielding a displacementof 4.92� 10�6 m3. To use the air compressor with refrigerant ratherthan air, the areas to be addressed included the requirement for apressurized crankcase, a modified intake manifold, crankshaft andthe overall structure of the compressor housing, and the addition ofa rotating seal.

The crankcase required pressurization to provide backsidepressure during startup. Without equalization of pressure on bothsides of the piston, the starting torque would be too high for thecrankshaft to be rotated. Backside pressure also helps equalizationof force on both sides of the piston, which leads to smootheroperation since the variation of torque on the crankshaft is not aslarge. With the presence of backside pressure on the piston from

Fig. 1. Air compressor details.

T.C. Ernst, S. Garimella / Applied Thermal Engineering 60 (2013) 316e324318

the refrigerant in the crankcase, the average pressure differential islower, minimizing the need for a larger flywheel for smoothoperation. The presence of refrigerant on the backside of the pistonalso alleviates the problem of refrigerant leaking past the pistonseal. Any refrigerant that leaks enters the crankcase, which isconnected to the low pressure side of the system, thus recapturingthe refrigerant. A Teflon rotating seal between the crankcasehousing and the input shaft was also added. The compressor head,valves and rotating seal were tested first. The structural integrity ofthe compressor was also evaluated to ensure that the crankshaft,connecting rod, piston and bearing surfaces were of sufficientstrength to enable compression of the refrigerant (which is 15times denser than air).

To meet the requirements for backside pressure and propersealing of the rotating seal, a vessel that completely enclosed thecompressor and head assembly was fabricated using a rectangularaluminum channel (Fig. 2). The back flange (e) was used to seal theend of the enclosure and also for an inlet port from the evaporator.The front flange (b) provided the front face of the rectangularenclosure for sealing purposes and also held the rotating seals andbearings for the shaft. The two endplates (b, e) were bolted togetherwith the rectangular channel (d) in between to form the enclosure.The additional bearing plate (a) was used to support the bearingsfor the rotational shafts. The seams were sealed using rubber gas-kets between the plates and the rectangular channel. Thecompressor was mounted to the aluminum plate and driven by theinput shaft passing through the plate.

Inside the compressor enclosure, polyol ester oil was used tolubricate the moving parts. The crankcase was located toward the

Fig. 2. Compressor e

bottom of the enclosure so that the crankshaft and connecting rodwere in contact with the oil at the bottom of the stroke. Thisallowed the crankshaft to lubricate the compressor by distributingthe oil throughout the enclosure. System power was supplied by a0.35 horsepower compressed air-driven motor capable of speedsranging from 0 to 10,000 RPM. The rotational speed of thecompressor was varied by adjusting the supply air pressure. Theoutput of the air-driven motor was geared in a 4:1 ratio to reducethe speed to that required by the compressor. The air motor wasconnected to the pinion by means of a flexible coupling and a6.4 mm diameter shaft. Power was transmitted to the input of thecompressor through the gear assembly (Fig. 3).

An aluminum microchannel, multi-louver fin heat exchanger(0.260 m wide, 0.238 m tall, 0.0211 m deep) was used as thecondenser. The condenser consisted of 24 rows of microchanneltubes with 615 fins per meter. This configuration offers highrefrigerant-side heat transfer coefficients, enhances air-side heattransfer due to the interrupted fins, and provides a large surface-to-volume ratio to yield a compact geometry. A shroud (fabricatedfrom acetate sheet) routed the air from the rectangular cross-section of the condenser to the circular cross-section of the axialflow fan. The refrigerant temperature and pressure were measuredat the condenser inlet and outlet to determine the thermodynamicstates. A thermostatic expansion valve with external equalizationexpanded the refrigerant to the low side pressure. This valveallowed for automatic regulation of the refrigerant flow based onthe superheat (3 �C) of the refrigerant leaving the evaporator,eliminating the need for manual control. The refrigerant tempera-ture and pressure were also measured at the evaporator outlet.

nclosure layout.

Fig. 3. Reduction gearing to compressor.

Fig. 4. Compressor test facility.

Fig. 5. Finalized portable cooling system.

T.C. Ernst, S. Garimella / Applied Thermal Engineering 60 (2013) 316e324 319

For preliminary testing purposes, an evaporator that allowedstraightforward measurement of the cooling duty was constructed.It consisted of a coiled tube-in-shell heat exchanger with coolant onthe shell side and refrigerant flowing inside eight aluminum tubes.Refrigerant was distributed in these 3.2 mm OD, 1.9 mm ID, 3.05 mlong tubes from a common header. Each tube was wrapped in aspiral pattern before returning to the exit header. The tubes wereplaced in an 8 L container of coolant used to measure the evapo-rator heat duty. A variable speed pump circulated the coolantduring testing. A cartridge heater provided the desired heat load tothe coolant. The coolant temperaturewasmeasured at the inlet andoutlet of the shell side of the evaporator. The temperature of theevaporator coolant was selected to provide a comfortable range forcooling of the body. The evaporator refrigerant-side temperaturewas slightly below the desired skin temperature to allow heattransfer while remaining comfortable for the user. The coolant flowrate (w1 LPM) was maintained high enough to prevent significanttemperature variation (<2 �C) between the inlet and outlet, whichprovided an almost constant temperature for the evaporator.

To simulate the elevated ambient conditions for a portablecooling system, the system (condenser and compressor assemblies)was surrounded in a 0.610 (W) � 0.381 (D) � 0.508 (H) m Plexiglasenclosure. Prior to testing, the air temperature was raised to37.7 �Ce43.3 �C inside the enclosure by means of a 1500 W airheater controlled by a thermostat. During testing, the heater wasnot used because heat rejected by the condenser maintained thedesired air temperature. Air flow across the condenser was pro-vided by an electric motor-driven axial flow fan (an electric motorallowed independent control of the fan speed). The fan blade was a0.178 m diameter high-speed axial flow propeller driven between3000 and 5000 RPM.

Testing was conducted over a range of ambient (37.7e47.5 �C)and coolant (26e30 �C) temperatures. Five refrigerant tempera-tures and pressures at the condenser inlet and outlet, the evapo-rator inlet and outlet, and the compressor inlet downstream of thesuction valve were recorded. Two thermocouples were used on thecoolant loop at the inlet and outlet to the evaporator shell-side and

two were used for the air flowing into and out of the condenser.Fig. 4 provides an overview of the test facility. The compressor wasable to deliver 330 W of cooling in an ambient temperature of37.7 �C. These tests demonstrated that the modified air compressorwas capable of operation under the required loads and flow con-ditions when used with refrigerant in the portable cooling systemunder consideration. After establishing feasibility, several modifi-cations (described below) were made to enable it to perform morereliably and efficiently, while being more compact and lightweight.

5. Overall system development

Upon conclusion of compressor testing, the final system wasdesigned and built to operate as a personal cooling system. Fig. 5shows the layout of the system arranged for use in a portablecooling system. Fig. 6 shows the components required for enginestarting, power generation, refrigerant compression and flow con-trol, air flow control and fuel storage. Toward the bottom center isthe engine, which provides input power to the system. On the rightis the compressor with the modifications to the intake manifoldand crankcase. The condenser at the back establishes the cross-sectional envelope of the system. Power to the compressor andthe condenser fan was delivered by means of a gear train (Fig. 7.)The input power to the cooling systemwas supplied by a Traxxas 2-stroke compression ignition engine with a displacement of 2.5 cm3

and rotational speeds of up to 30,000 RPM. It delivered power tothe gear train by means of a centrifugal clutch, which allowed the

Fig. 6. Isometric views of system.

T.C. Ernst, S. Garimella / Applied Thermal Engineering 60 (2013) 316e324320

engine to be started independent of the gear train and the load ofthe compressor and fan. After the engine was started, the throttlewas increased, causing the centrifugal clutch to engage and rotatethe drive train. The drive train was composed of three parallel andone perpendicular shafts (Fig. 7.) The input was first geared down ina 3.5:1 ratio from the engine to a shaft that was subsequentlygeared down in a 4.4:1 ratio; this low speed shaft rotated thecompressor. This resulted in a speed reduction from the engine tothe compressor of 15.4:1. From the compressor shaft speed, agearing ratio of 1:4.4 was used to increase the speed for thecondenser fan. Two bevel gears were used in a 1:1 ratio to changethe axis of rotation for the condenser fan by 90�. The gear train washeld in place by two plates that held high speed bearings and werepart of the integrated structure supporting the system. The enginewas run on a mixture of methane (78%), nitro-methane (10%), andcastor oil (12%). As it was a two-stroke engine, the fuel containedthe lubrication oil (castor oil) for the moving components. The fueltank is slightly pressurized by the engine exhaust to maintain fuelflow to the engine, with the flexible hose connecting the mufflerand fuel tank (Fig. 5.)

The overall system consisted of many of the same componentsthat were used for the compressor testing, with modificationswhere necessary. The condenser remained unchanged except forchanges in shroud geometry for air flow. The enclosure for thecompressor that served as a reservoir of refrigerant for the intakevalve was eliminated. Back pressure was provided to the piston bysealing the crankcase using an aluminum plug clamped on the

Fig. 7. Reduction gear train.

opening of the crankcase as shown in Fig. 8. This plug allowed thecrankcase to be extended outward for additional volume forlubricating oil as well as refrigerant. The compressor shaft traveledthrough a brass hub which held the two bearings and two Teflonseals that prevented leakage during rotation of the shaft. Refrig-erant was delivered to the intake valve using a cylindrical plenumsealed to the compressor head. The plenum had ports for a pressuregauge, refrigerant flow from the evaporator, and crankcase pres-surization. This last tube connected to the crankcase plug, enablingthe piston to have the required backside pressure for properoperation. These modifications reduced weight and size, addedsimplicity, and increased the ability for heat rejection to the sur-rounding air directly from the compressor. Rather than sealing theentire volume surrounding the compressor, the crankcase wassealed and the compressor body itself became the pressure vessel,which reduced weight and size considerably. With the redesignedsystem, the compressor itself was exposed to air, which enabledincreased heat removal.

For the testing of the portable cooling system, the engine wasused as the source of power for both the fan and the reconfiguredcompressor. The entire systemwas placed inside a chamber (Fig. 9)that provided air at controlled elevated temperatures. An enclosureensured that condenser air flow was generated solely by the fanintegral to the cooling system. The air was drawn in the face of thecondenser and discharged across the compressor and engine,which allowed both components to be cooled by the passing air.Engine speed was controlled by a throttle linkage that extendedoutside the enclosure. The fuel tank was also located outside toenable refilling. The engine exhaust was routed to a fume hood.Electrical connections were also routed to the exterior to allow thebattery pack to be connected to the starting circuit of the engine.The majority of the cooling system components were containedinside the chamber since the entire systemwould be at the elevatedambient temperature during use. However, for the control of theevaporator heat duty, refrigerant lines were routed out of thechamber to the evaporator. A needle valve served as the expansiondevice, and was located outside the chamber to allow adjustmentsduring testing and control of the low-side pressure. Transparenthose upstream of the expansion valve enabled verification of sys-tem charge and the subcooled state at the condenser exit. The lines

Fig. 8. Compressor layout.

Fig. 9. Test facility schematic.

T.C. Ernst, S. Garimella / Applied Thermal Engineering 60 (2013) 316e324 321

to and from the evaporator were insulated to minimize unwantedheat transfer.

A 50/50 ethylene glycol/water solution served as the evaporatorcoupling fluid. The heat duty of the evaporator was obtained bymeasuring the voltage and current to the cartridge heater using adigital voltmeter and digital ammeter. The voltage to the heaterwas varied using a variable transformer. The evaporator heat dutycomputed from the coolant temperature difference and flow ratewas used to validate this heater input. A rotameter was used tomeasure coolant flow, which was provided by a variable speedpump. Inside the 8 L coolant tank, the refrigerant tubing had thesame spirally wound pattern described earlier, with the coolantflowing in a counter flow arrangement to the refrigerant. Refrig-erant from the evaporator flowed through a transparent hose toenable checking that it was superheated. It then returned to thesuction side of the compressor inside the chamber via anotherhose.

As shown in Fig. 10, temperatures were measured using type Tthermocouples at 6 refrigerant locations, 3 glycol/water locations,and 3 air locations. The high and low side refrigerant pressureswere also measured using pressure gauges. In addition, coolant

Fig. 10. Cooling system performance summary.

flow rate was measured along with the heater voltage and current.Engine speed was measured using a digital stroboscope. Testingwas performed over an ambient temperature range of 37.7e47.5 �C,evaporator temperature range of 22.2e26.1 �C, and engine speedsof 10,500e13,300 RPM, with three values for each parameteryielding a total of 27 points.

6. Data analysis

The analysis of the data from these tests is explained here using adata point at the following conditions for illustrative purposes:ambient temperature ¼ 43.5 �C, evaporator average coolanttemperature ¼ 29.9 �C, and engine speed ¼ 13,300 RPM. Heatervoltage and amperage measurements yielded a heater input of225 W. A small portion of this heat load is attributed to thermalstorage in the loop during the test due to slight variations in thecoolant tank temperature if the heat input did not exactly match theevaporator heat duty. With a coolant mass of 8.55 kg, and time rateof change of temperature of �1.88 � 10�4 �C s�1 for this case, thethermal storage term is only �5.5 W (negative values indicatetransient tank cooling instead of heating), compared to the heaterinput of 225W, yielding an evaporator heat duty of 230W. Over therange of experiments in this study, the thermal storage portion ofthe evaporator load varied from 1 to 12%. The refrigerant flow rate iscalculated from the evaporator heat duty and the refrigerant inletand outlet enthalpies based on measured temperatures, pressuresand qualities (as appropriate) using Engineering Equation Solver(EES) Software [9]. For this case, the refrigerant flow rate was1.60 � 10�3 kg s�1. With the refrigerant mass flow rate known, thecondenser heat rejection based on the refrigerant side was calcu-lated. At the measured condenser pressure of 1370 kPa (saturationtemperature of 51.5 �C), from the corresponding superheated inletand subcooled outlet temperatures and enthalpies, and the refrig-erant mass flow rate reported above, the condenser heat rejectionwas 275 W. Similarly, for low and high side pressures of 663.8 and1370 kPa, respectively, the compressor work is 46W, resulting in an

T.C. Ernst, S. Garimella / Applied Thermal Engineering 60 (2013) 316e324322

estimated isentropic efficiency of 0.527. The refrigerant-sidecondenser heat duty was compared with the corresponding air-side value. Fan speed measurements using a stroboscope and airvelocities measured with a turbine anemometer were correlated toyield face velocities and condenser air flow rates. These values wereused in conjunction with the air outlet and inlet temperatures toobtain air-side duties, which, across the test matrix in this study,were on an average, different from the refrigerant-side duties by 9%.Some of this difference is attributed to variations in the air tem-perature across the face of the condenser. Similarly, for the evapo-rator, the average difference between the heater measurements andthe coupling fluid-side heat duties across the test matrix was 5%.

With an evaporator heat duty of 230 W and a compressor workinput of 46 W, the refrigeration cycle coefficient of performance(neglecting fan work) is 5.0. The engine fuel consists of 78%methanol, 10% nitro-methane and 12% lubricating oil [15]. Withnitro-methane modeled as methanol [3], and an average fuel flowrate of 0.316 kg h�1, the energy consumption rate is 1749 W. Thecooling duty of 230 W yields a system efficiency of 0.132. Heatlosses and gains between the system and the ambient werecomputed to obtain an estimate of the parasitic load. The compo-nents most likely to affect performance include tubes connectingthe components in the elevated temperature environment or theroom temperature environment, and the evaporator tank. Withinsulated components, and small temperature differences betweenthe components and the environment, the natural convection co-efficient is onlyw1.6Wm�2 K�1, and after accounting for radiation,the net heat loss from the systemwas estimated to bew2W, whichis negligible compared to a cooling duty of 230 W.

7. Results and discussion

Tests were conducted at three ambient temperatures (37.7, 43.3and 47.5 �C), three evaporator coolant temperatures (26, 28 and30 �C), and three engine speeds (10,500, 12,250 and 13,300 RPM).Fig. 10 displays the results of the entire test matrix; the influence ofindividual parameters is discussed below.

At constant ambient temperature, the heat duty of the evapo-rator increases with increasing engine speed due to the largerrefrigerant flow rates. At the highest engine speed of 13,300 RPM,the evaporator heat duty was 280 W when tested at the highestcoolant temperature (30 �C) and lowest ambient temperature(37.7 �C). The cooling duty decreases when the evaporator coolanttemperature is lowered because the temperature lift between thelow side and high side is increased. The larger the temperature lift,the higher the pressure ratio is in the compressor and the lower theperformance of the system. The pressure ratio for the least extremecase (evaporator coolant temperature of 30 �C, and ambient tem-perature of 37.7 �C) was 1.7, whereas for the most extreme case(evaporator coolant temperature of 26 �C, and ambient tempera-ture of 47.5 �C) the pressure ratio was 2.3. As the ambient tem-perature increased from 37.7 �C to 47.5 �C, the condenser saturationtemperature increased from 44.2 �C to 51.5 �C, with a corre-sponding pressure increase from 1136 kPa to 1370 kPa. With higherpressures, the compressor input work is increased, which leads toadditional heat rejection. Finally, with higher ambient tempera-tures, the system would also have higher heat gains to the evapo-rator from the surroundings. Therefore, the increased ambienttemperature has a negative effect on the cycle in multiple ways. Theincreased compression for a given input power reduces therefrigerant mass flow rates, and therefore the cooling load. For thetwo cases mentioned above, the refrigerant mass flow rate de-creases from 0.0015 kg s�1 for the least extreme case to0.0010 kg s�1 for the most extreme case for an engine speed of12,250 RPM.

It should be noted that when engine speed increases, thecondenser fan speed also increases proportionately due to thedirect coupling, leading to increases in condenser heat rejection,refrigerant mass flow, and cycle performance. Compressor effi-ciency might also be higher at the higher speeds. The effect ofcoolant on evaporator heat duty was much less pronounced at thehigher ambient temperatures (47.5 �C). This is because the variationin coolant temperature is much smaller relative to the overalltemperature lift between the ambient and the coolant for this casethan in the lower ambient temperature cases. The temperature liftsbetween the ambient air and evaporator coolant were 17.5,19.5 and21.5 �C for the three cases shown. Refrigerant temperature in theevaporator was typically 3e4 �C lower than that of the coolant.

At a constant engine speed of 13,300 RPM, the evaporator heatduty decreases rapidly with the increase in ambient temperature.This is most notable in the case with the highest engine speed.Increased ambient temperature is detrimental to the performanceof the system due to the increased difficulty of rejecting heat to thesurroundings, and the corresponding increased compressor work.The duties are lower at the intermediate engine speed; however,they do not drop off as rapidly at higher ambient temperature as inthe case of 13,300 RPM. This is because evaporator heat duty isproportionately lower at the decreased engine speeds. The refrig-erant mass flow rate, and therefore evaporator heat duty, isdependent on the compressor speed. With proportionately lowerduties, the actual change in duty is smaller. The difference in heatduty due to different values of ambient temperature is less pro-nounced at the lower values of coolant temperatures. This is due tothe fact that as the overall temperature difference between the lowand high sides increases, the individual effect of the variation ofambient temperature by itself becomes less significant.

Limited testing was also conducted at higher engine speeds. Thedrawbacks of the higher engine speeds are an increased fuel con-sumption rate, more severe loading on the overall system, andincreased noise. For these data points, the other variables were heldconstant at nominal design values (evaporator coolant temperatureof 28 �C, ambient temperature of 43.3 �C). The evaporator dutycontinued to increase as engine speed was increased beyond theoriginally chosen range, however, the rate of increase is lower. Thesystem was designed for portability (light-weight and compact);therefore, higher engine speeds, which raise strength and dura-bility concerns, were not tested extensively.

The amount of fuel used was measured for series of testsgrouped in sets. Testing was performed in three stages where ninedata points each were taken. This was due to the fact that the testchamber took an appreciable length of time to stabilize at thedesired ambient temperature. During testing, the fuel consumptionof the entire set of data for each ambient temperature was recor-ded. This gave an overall average for the fuel usage for each ambienttemperature. Fuel consumption increased with increasing ambienttemperature to achieve the additional compression. As reportedabove, for the representative test point, the efficiency of convertingfuel energy into evaporator cooling was 13.2%. Although this effi-ciency is relatively low, the engine-driven system still has a higherenergy density than the other methods examined, such as power-ing with batteries or a fuel cell. Small scale engines are typicallyinefficient at converting fuel energy to usable shaft work. The fuelconsumption at ambient temperatures of 37.7, 43.3, and 47.5 �C was0.269, 0.316, and 0.340 kg h�1, resulting in system efficiencies of15.2, 10.2 and 8.0%, respectively. With increased ambient temper-ature, not only does the fuel flow rate increase, but the evaporatorheat duty also decreases. This would cause an increased fuel storagerequirement and an associated increase in system weight to pro-vide cooling for the desired duration at elevated temperatures.With the use of an integrated 2 L fuel tank, the system can provide

Fig. 11. Wearable evaporator.

Table 1System mass.

System/component Mass

Refrigeration system 1.76 kgCompressor 0.44 kgCondenser 1.22 kgExpansion valve 0.10 kg

Wearable evaporator 0.85 kgInsulating vest 0.32 kgAluminum foil/tubes 0.53 kg

Power supply system 1.57 kgEngine and base 0.71 kgBattery pack 0.43 kgGears 0.43 kg

Support system 1.13 kgBackpack 0.26 kgStructure/hardware 0.87 kg

Overall total 5.31 kg

T.C. Ernst, S. Garimella / Applied Thermal Engineering 60 (2013) 316e324 323

cooling for the three ambient conditions (37.7 �C, 43.3 �C, 47.5 �C)for 6.7, 5.7 and 5.3 h, respectively, while cooling at a rate of 226, 178and 149 W, respectively. For the increased speed tests, with evap-orator coolant temperature of 28 �C and an ambient temperature of43.3 �C, the average fuel consumptionwas 0.358 kg h�1 over enginespeeds from 14,000 to 19,250 RPM and an average evaporator heatduty of 268 W. These cooling duties and the durations are muchhigher than those reported in the literature for small-scale coolingsystems.

8. Wearable system

Once the performance of the system described above wascharacterized using a simulated heat load supplied from a heaterand a fluid tank, an evaporator incorporated into a garment worn incontact with the human body was developed. The wearable evap-orator was first designed using heat transfer analyses to determinethe appropriate geometry. An actual cooling garment was thenstitched and assembled in-house to test it in the form of a wearablecooling system operating in an elevated temperature environment.

The cooling garment model consisted of three layers includingthe insulation, the heat transfer surface (aluminum foil) and the

Fig. 12. Wearable cooli

refrigerant tube layer. The refrigerant tube side would be worntoward the body, with the insulation layer shielding the cooledinterior from the surrounding ambient. Tubing attached to a thinaluminum foil layer was used. The tubing carries the refrigerant asit removes heat from the body while evaporating. The foil layerserves as a fin to remove heat from a larger area and transfer theheat energy to the refrigerant tubes. Standard bare and finnedsurface heat transfer calculations were conducted to size the tubediameter, pitch, and foil thickness. The analysis was first performedassuming the garment encompassed the entire body. This yieldedestimates of the required geometrical features such as number,pitch, and diameter of tubes, foil thickness, surface area, andinsulation thickness.

The actual wearable evaporator that was stitched and assem-bled covered only the torso, which, as discussed earlier, is the mosteffective area to cool [5], since it has the largest surface area andgenerates a significant amount of heat. The cooling vest wasassembled using a neoprene wetsuit (Fig. 11). The 3 mm thickwetsuit served as the insulating outer shell of the cooling garmentto minimize heat gain from the surroundings. The tubing OD was3.2 mm, with an ID of 1.9 mm, the same as the test evaporator. Thealuminum foil was 0.0508 mwide and had a thickness of 0.15 mm.Each tubewas centered on a strip of foil. Thin fabric mesh sewn intothe neoprene shell held the strip of foil and tube in place inside the

ng system testing.

T.C. Ernst, S. Garimella / Applied Thermal Engineering 60 (2013) 316e324324

shell. The foil had adhesive on the side toward the tube which heldthe two bonded together. The garment had a total of four tubingpasses in parallel sewn into it, each supplied by a flexible hose fromthe backpack structure. The individual tubes were 2.44 m in length,for a total length of tubing in the garment of 9.8 m and a surfacearea of 0.495 m2. Each of the four passes of tubing traversed up anddown twice inside the garment at a 51 mm pitch before beingrouted back to the backpack structure in another flexible tube.Testing was performed using the wearable evaporator at anambient temperature of 40.5 �C provided by a controlled environ-ment chamber. The system operated satisfactorily by providingcooling to the user while jogging on a treadmill at 7.2 km h�1. Fig.12displays the cooling garment and portable cooling system as wornby the user.

The individual masses of each component, as well as the overallsystem (excluding the fuel) are shown in Table 1. For a volume of2 L, the resulting fuel mass would be 1.8 kg.

9. Conclusions

A wearable personal cooling system was designed, fabricatedand tested in this study. A vapor compression system was chosenbased on criteria such as simplicity, performance, and the numberof components. Similarly, a liquid fuel-based small-scale enginewas chosen based on a higher overall energy density from amongoptions such as fuel cells and lithium-ion batteries to supply powerto the vapor compression cooling system. The engine powered thevapor-compression system compressor and the condenser fanthrough appropriate gear trains. Several modifications to an off-the-shelf portable air compressor were designed and fabricatedin-house to convert it into a refrigerant compressor required forthis application. Detailed tests on a stand-alone compressor teststand were first performed to establish the viability of thiscompressor for refrigerant compression. The compressor was in-tegrated into the overall cooling system and system performancewas investigated over a wide range of controlled, elevated ambienttemperatures, evaporator temperatures and engine speeds. Thecooling system was successful at providing cooling at a levelequivalent to the heat produced by a typical subject working at alevel comparable to calisthenics or moderate exercise at elevatedambient temperatures. Thus, the system demonstrated heatremoval rates between 100 and 300 W at ambient temperatures inthe range 37.7e47.5 �C, engine speeds of 10,500 to 13,300 RPM, andevaporator coolant temperatures of 26e30 �C. Fuel consumptionrates varied from 0.269 to 0.340 kg h�1 for this same range ofconditions. The cooling duty increased at higher engine speeds, butdecreased at higher ambient temperatures and lower evaporator

temperatures. Additional testing at higher engine speeds(13,300 < RPM < 19,250) showed that the resulting increase incompressor speed leads to higher cooling capacities.

The backpack mounted wearable cooling system developed inthis study had a total mass of 5.31 kg and can provide cooling at therate of about 178 W for a duration of 5.7 h at a nominal ambienttemperature of 43.3 �C. This eliminates the need to be tethered to asource of input power or cooling fluid, which some systemscurrently employ for personal cooling. Thus, the system will helpreduce the effects of heat exhaustion and fatigue, thereby reducingthe stress level on the body, while increasing productivity andsafety. This system is expected to benefit hazardous duty personnelsuch as firefighters, military, or factory personnel who work atelevated temperatures. In actual use, care should be taken toconsider the suitability of the use of fuel-driven systems in haz-ardous environments, and the potential effects of chemical andparticulate matter on system performance.

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