Erik Björk
email: bjork@energy.kth.se

Björn Palm

Johan Nordenberg

A thermographic study of the on-off behavior of an all-refrigerator

Keywords: household refrigerator, thermography, heat exchanger, cooling system

In this work a thermographic camera is used to observe the temperature distribution of a household refrigerator cooling system operating at on-off cycling conditions. This technique offers an alternative method to analyze the cooling system compared to conventional thermocouples. In particular it is interesting to view the overall picture of how the refrigerant charge is distributed over the cooling system at transient conditions. In additions, four sources of energy losses were indentified and discussed.

Out of these losses, two would have been difficult to find using conventional thermocouples temperature measurements.

Introduction

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The global annual production of household refrigerators and freezers is more than 80 million units. With an expected lifetime of 10 to 20 years the number of units in use is enormous and therefore their impact on the global energy consumption significant. It follows that domestic refrigeration is a natural target for energy consumption controls in the EU and in many countries around the world.

Most studies that have expanded the knowledge of the working principle of small, onoff controlled, capillary tube throttled, vapor compression cooling systems (i.e. household refrigerators & freezers) are based on thermocouple recordings. Useful experimental data have for instance been provided by Wang and Wu [START_REF] Wang | Start up and shut down operation in a reciprocating compressor refrigeration system with capillary tubes[END_REF], Janssen et al. [START_REF] Janssen | Cycling losses in domestic appliances: an experimental and theoretical analysis[END_REF], Rubas and Bullard [START_REF] Rubas | Factors contributing to refrigerator cycling losses[END_REF], Jakobsen [START_REF] Jakobsen | Energy Optimisation of Refrigeration Systems[END_REF], Krause and Bullard [START_REF] Krause | Cycling and quasi-steady behaviour of a refrigerator[END_REF], Coulter and Bullard [START_REF] Coulter | An experimental analysis of cycling losses in domestic refrigerator-freezers[END_REF] and Björk and Palm [START_REF] Björk | Performance of a domestic refrigerator under influence of varied expansion device capacity, refrigerant charge and ambient temperature[END_REF]. Experimental charge inventories were reported by Kuijpers et al. [START_REF] Kuijpers | Experimental verification of liquid hold-up predictions in small refrigeration heat exchangers[END_REF][START_REF] Kuijpers | Refrigerant charge in small appliences; Influence of the capillary / suction line heat exchanger[END_REF], Mulroy and Didion [START_REF] Mulroy | Refrigerant migration in a split-unit air conditioner[END_REF], Björk and Palm 2006 [3-4]. X-ray was used by Inan et al. [START_REF] Inan | X-ray investigation of a domestic refrigerator. Observations at 25°C ambient temperature[END_REF] to demonstrate the refrigerant movement in selected parts of a household freezer cooling system. Björk [START_REF] Björk | Slug formation in a domestic refrigerator with plate evaporator[END_REF] visualized the refrigerant flow in a household refrigerator using a transparent glass lid evaporator.

Evidence of using thermography in household refrigeration is sparse. Cristalli et al., [START_REF] Cristalli | A fast on line quality control of refrigerators based on thermal image detection and power consumption[END_REF] proposed an online method based on thermography to identify manufacturing and assembling defects during production of household refrigerators. Herolf [START_REF] Herolf | Free convection heat and mass transfer in a domestic refrigerator[END_REF] used an infrared camera to visualize the air temperatures in a vertical cross section of the cabinet.

In this study the results from an unusual experimental run is presented. A household refrigerator cooling system is observed using an infrared video camera during normal They are characterized by low cooling capacities (50-250 W), low quantities of refrigerant charge (20-200 g), a refrigerant accumulator located at the evaporator outlet, a hermetically sealed cooling system, a capillary tube expansion device which is in heat exchange with the suction line, and a cooling capacity control by on-off control.

The on-off cycling

The most common way to control the cooling capacity in household refrigerators and freezers is by on-off control (cycling or intermittent operation). The compressor starts as a preset temperature in the refrigerated space is exceeded (cut-in temperature) and shuts down as a low temperature (cut-out temperature) is reached. A typical course of events at compressor start and stop is as follows: At the moment before the compressor starts, the evaporator is at its highest temperature. Most refrigerant, by mass, is located in the evaporator and compressor whereas the condenser and the capillary tube contain only superheated gas. As the compressor starts, the initial compressor mass flow is high (high evaporation pressure) whereas the capillary tube mass flow is low (superheated gas). The result is that refrigerant is displaced towards the condenser which may starve the evaporator with refrigerant liquid. As a result the

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evaporator pressure decreases thereby lowering the compressor mass flow. A liquid pool is formed at the condenser outlet which increases the mass flow through the capillary tube. Now the evaporator "refilling" begins. This means that the dry-out point (position in evaporator where superheat starts) moves downstream gradually 'activating' the evaporator as the wetted area, active for evaporation, increases. It is important that this process is short in comparison to the compressor on-period since the system during this time period operates inefficiently due to the superheat and lowered evaporation temperature.

When the cut-out temperature is reached the compressor stops and the pressure equalizes within the system as refrigerant continues to flow through the capillary tube.

The condenser pressure decreases and slowly approaches the saturation pressure in the evaporator while refrigerant is forced into the evaporator. During this migration process it is important that the refrigerant liquid is freely drained towards the condenser outlet. If pockets of liquid remain in the condenser as the capillary tube inlet liquid seal is broken a reversed heat transfer follows, where liquid in the condenser evaporates and condensation occurs in the evaporator.

The charge displacements at startup and shut-down affect the performance negatively.

Totally, efficiency losses of 5-37% have been reported [START_REF] Björk | Refrigerant mass charge distribution in a domestic refrigerator. Part I. Transient conditions[END_REF][START_REF] Coulter | An experimental analysis of cycling losses in domestic refrigerator-freezers[END_REF][START_REF] Jakobsen | Energy Optimisation of Refrigeration Systems[END_REF][START_REF] Janssen | Cycling losses in domestic appliances: an experimental and theoretical analysis[END_REF].

Methods

A single compartment refrigerator with natural convection (passive) heat exchangers was studied using thermocouples and an infrared camera. At steady cycling conditions

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(each on-off period at similar length) the variations in the surface temperatures of the evaporator and condenser were recorded. The temperatures of the lower half part of the condenser were also separately recorded from close range to give a more detailed picture. The experiments were carried out in the following way: At 25°C ambient temperature and the refrigerator thermostat set to warmest possible (averaged air temperature in the cabinet 7.2 °C) the refrigerator was running until it was steadily cycling. Then the evaporator was filmed during one on-off period. Next, the video camera was repositioned to capture one on-off period of the whole condenser and thereafter the lower part of the condenser. Finally, the three video recordings (evaporator, overall condenser and condenser close-up) were synchronised so that images of the evaporator and condenser could be viewed side by side displaying the same moment in the on-off period. This was done by synchronising the condenser films to the evaporator film. They were synchronised at two moments in the on-off period; the moment when the compressor started, and the moment when the compressor stopped. This means that the images showing the evaporator and condenser comes from similar, but not the same, on-off period. However, the temperature recordings and the evaporator films come from the same on-off period. Figure 1 shows an example of the synchronised videos.

Test object

The test object was an Electrolux ER8893C, single door and single compartment refrigerator with the following data (see also Figure 2

Modifications to test object

The IR-camera used was an AGEMA Thermovision 400 with a 40 mm/40° lens. To avoid thermal reflections influencing the IR camera, the evaporator and condenser were painted dull black. This modification did not change the refrigerator behaviour in any noticeable way. This type of refrigerators that has natural convection condensers located at the back of the refrigerator is optimised in terms of refrigerant charge and capillary tube capacity for a condition where they are placed against a wall [START_REF] Björk | Performance of a domestic refrigerator under influence of varied expansion device capacity, refrigerant charge and ambient temperature[END_REF]. The reason is that the "chimney" effect enhances the air flow across the condenser in this position.

However, this would make IR recording of the condenser difficult. Instead, during the experiments a thin transparent plastic film, of the type used for food storage, was stretched over the cabinet's back side forming a channel with geometry equal to that formed when the refrigerator is positioned against a wall. Through this, a chimney for the condenser heat was formed while at the same time IR recording of the condenser was enabled. It was experimentally verified that the system behaviour was similar to the wall positioned refrigerator. The error by filming the condenser through the plastic film was measured to be less than 0.1 K (by quickly removing and inserting the plastic film).

Due to the rather narrow angle 40° lens, the camera needed to be at some distance to capture the entire evaporator. Therefore an inspection opening was made in the refrigerator door and some insulation material was used to locate the camera at the

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8 needed distance from the evaporator. This is why the insulation material sticks out somewhat from the door (Figure 3). Even though this opening was as much as possible insulated, an increased heat leakage was unavoidable. The result of this was a prolonged cycle length in comparison to the unmodified system. The on-period was extended from about 700 s to about 1000 s, the cycle length from about 2700 s to about 3200 s.

The experimental set-up with the IR-camera pointing through the inspection opening in the door can be seen in Figure 3 and Figure 4. Temperatures were measured by thermocouples located as seen in Figure 2.

Uncertainties

The estimated uncertainties in temperature were 0.5 K (U=95%) for the thermocouples and 5 K (U=95%) for the IR camera. The difference in time was less than 2% (U=95%) of the different on-off periods. Since the condenser films were synchronised to the evaporator films at the moment when the compressor started and stopped, the largest uncertainty between the two films would be at the end of the on and off-period. In the presented images, the largest uncertainty will occur at 720 s.

The uncertainty in time at this moment is about 20 s (U=95%) between the evaporator

Results

Firstly some comments will be made about the thermocouple temperature recordings that, as was mentioned before, are taken from the same on-off cycle as was used for the IR film of the evaporator. After this the thermo graphic images will be presented and discussed.

Temperature recordings

In Figure 5 the temperatures during one on-off cycle are shown. It is seen that the cycle length is more than 3200 s (3240 s). It is also seen that the on-period is approximately 1000 s (994 s). Following a startup, the suction line temperature (T8) immediately falls from 25°C to below 10 °C, after which it increases to above the ambient temperature. This temperature dip is caused by liquid refrigerant, accumulated in the evaporator during the off-period, is being drawn into the suction line where it evaporates on its way towards the compressor. This liquid suction, occurring at compressor startup, was previously observed by Rubas and Bullard [START_REF] Rubas | Factors contributing to refrigerator cycling losses[END_REF], Krause et al. [START_REF] Krause | Cycling and quasi-steady behaviour of a refrigerator[END_REF] and Björk and Palm [START_REF] Björk | Refrigerant mass charge distribution in a domestic refrigerator. Part I. Transient conditions[END_REF]. It is explained by the initially "overcharged" evaporator and the initially high On the evaporator side the inlet temperature (T5) falls sharply at startup, whereas the accumulator inlet temperature (T6) is at a higher temperature until almost 700 s when they coincide. This is the time needed for the refrigerant liquid to reach the accumulator inlet thus "activating" the evaporator after the initial starving at the compressor startup. In other words, this is the time needed for the dry-out point to move the distance from the evaporator inlet to the accumulator inlet. The evaporator outlet temperature (T7) is at a higher temperature than T5 and T6 throughout the onperiod. This superheat is almost 3 K at the moment when the compressor stops.

In the off-period the condenser temperatures (T2-T4) asymptotically approaches the ambient temperature. The evaporator temperatures (T5-T7) first increase to 0 °C, where they level out while the evaporator defrosts. Then they increase again to slowly approach the cabinet air temperature. It is seen that the defrost time at the T7 location is shorter compared to T5 and T6. This may be explained with less frost at this location. The temperature at this location is higher during the on-period which means

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11 that less frost is formed here. It can be mentioned that the frost layer was very thin after the on-period (estimated by visual inspection to be < 0.5 mm).

The compressor temperature (T1) varies between 35 and 45 °C during the complete on-off cycle. It is worth noting, that this thermally heavy component will always stay significantly above the ambient temperature. The heat dissipated from the compressor shell is used to evaporate the defrost water collected in a plastic container on top of the compressor (see also Figure 4). The time to "activate" the evaporator in the on-period was unexpectedly long. The typical activation time for this kind of refrigerator is less than 180 s. In this case it was about 700 s. The explanations are the high temperature setting in the refrigerator and the increased heat load due to the modified door (inspection opening). Together these two factors give a higher evaporation temperature than under normal operation, thus resulting in an increase of the compressor mass flow. Since the capillary tube has a fixed flow resistance, longer time will be needed to activate the evaporator at startup.

Thermographic images

Images from the IR video recordings are displayed in Appendix A. On the right side of the images the time is shown referring to the same time scale as in Figure 5. For instance, 0 s is the time when the compressor starts (frame picked out at the moment before any thermal change due to the compressor start can be seen). 994 s is the time when the compressor stops (frame picked out at the moment before any thermal change due to the compressor stop can be seen).

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At 0 s the evaporator is isothermal (within the resolution of the measurements). The condenser is also more or less isothermal. The compressor can be seen as a warm object at the bottom of the condenser images. Since this image shows the temperatures at the moment when the compressor starts but just before any thermal change is visible the compressor heat stems from the previous on-period.

At 10 s it is seen that cooling occurs in parts of the evaporator. One can identify the location of the accumulator and the vertical tube at the left side of the evaporator.

Obviously this is liquid which is accumulated in the evaporator during the off-period that evaporates as soon as the compressor starts. Apparently, part of this liquid was accumulated in the bottom part of the evaporator and most likely the rest in the accumulator itself.

The suction line, which is seen left of the compressor, is also rapidly cooled. Taking into account that the refrigerant exiting the evaporator has to pass through a two meter long coaxial type suction line heat exchanger (SLHX) -at ambient temperaturebefore reaching this location it is obvious that some refrigerant liquid is exiting the evaporator. In fact, in a previous work by Björk and Palm [START_REF] Björk | Refrigerant mass charge distribution in a domestic refrigerator. Part I. Transient conditions[END_REF] it was found from experiments using a refrigerator at similar conditions that about 15 g of refrigerant, or about 40% of the entire charge, was exiting the evaporator during the first 20 s following a compressor start. In the same article it was calculated that a cooling power of 1.7 kJ/cycle was uncounted for when performing an energy balance over the evaporator and SLHX at startup. A cooling power of 1. Another observation from the image at 10s is that the compressor discharge pipe is rapidly heated.

At 20 s refrigerant has made the full loop from the evaporator, through the condenser and back to the evaporator again. This can be seen at the evaporator inlet which cools down rapidly. From now on it is possible to see the dry-out point moving downstream gradually activating the evaporator. The accumulator is cold but not as cold as the evaporator inlet. Furthermore, in the following images (30 and 60 s) the accumulator will heat up. Apparently all liquid remaining in the accumulator from the off-period, has already exited the evaporator.

At 60 s sub-cooling is seen in the condenser approximately covering the four lowest horizontal pipes. From 60 to 180 s this sub cooling area decreases. At the same time the dry-out point advances forward in the evaporator. Apparently liquid is being transferred from this sub-cooled part of the condenser to the evaporator. After 180 s the sub-cooled area is no longer changing, and covering approximately the lowest two horizontal pipes in the condenser. This is in agreement with the T.C. temperature recordings (Figure 5) which revealed a constant sub-cooling in the later on-period.

Apparently, the liquid needed to further refill the evaporator is not taken from the condenser. Previous studies [START_REF] Björk | Refrigerant mass charge distribution in a domestic refrigerator. Part I. Transient conditions[END_REF][START_REF] Janssen | Cycling losses in domestic appliances: an experimental and theoretical analysis[END_REF] have shown that this liquid comes from the compressor oil, where it is dissolved during the off-period, and from the low side vapor whose pressure (density) decreases with the decreasing evaporator temperature.

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At 720 s the accumulator can still be seen as an area of somewhat higher temperature than the rest of the evaporator. At this time the T.C. temperature recordings (Figure 5)

show that the dry-out point has moved into the lower part of the accumulator.

At 994 s, which is the moment when the compressor stops, the evaporator is more or less isothermal. A vague color difference indicates the accumulator which at this time has a 3 K superheat at its outlet (Figure 5). It is possible that the temperature range (lower limit set to -15.8 °C) was selected too high making this superheat difficult to spot. It is also seen that the evaporator edges on the lower left side are at a higher temperature. Apparently the refrigerant channels (see Figure 2) are too distant from the edges. The result is that the evaporator plate is poorly used for heat transfer which leads to lowered system efficiency.

At 1054 s, which is 60 s after the compressor has stopped, the heat exchanger temperatures have changed significantly. It is interesting to note that there is no local heating at the evaporator inlet as a result of refrigerant migrating from the condenser into the evaporator. Intuitively it may feel like "hot" refrigerant liquid coming from the condenser would cause a heating effect. However, any liquid passing the capillary tube will depressurize to the evaporator pressure and enter as a two-phase mixture at the saturation temperature corresponding to the pressure in the evaporator. The result is that the entering vapor increases the evaporator pressure so that some condensation occurs in the evaporator. This heating will not affect the evaporator locally at the evaporator inlet but rather globally all over the evaporator where condensation occurs.

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At 1114 s the image shows a higher temperature, evenly distributed, where the refrigerant lines are located. This can be attributed to heating by condensing refrigerant in the channels, as was explained above.

At 1134 s (the last image) a design weakness is revealed in the mid picture. Even though the condenser design admits a free fall towards the condenser outlet to avoid liquid pockets during the off-period, the filter dryer (lower left side) is at a significantly lower temperature than other parts of the condenser. Apparently some liquid is trapped in this device where it is evaporating during the off-period. It was estimated, based on the IR-films and the physical size of the filter dryer, that less than 0.3 kJ of heat was transferred to the filter dryer during the ten minutes period when this cooling effect existed. This heat will add to the cabinet thermal load in the following way. When the liquid seal at the capillary tube inlet is broken so that the capillary tube only contains vapor, any liquid left in the condenser will evaporate and, just like in a thermo-siphon, the corresponding heat of evaporation will be released during condensation at the coldest point in the evaporator. This thermal transport is obviously an energy loss. In previous articles [START_REF] Björk | Refrigerant mass charge distribution in a domestic refrigerator. Part I. Transient conditions[END_REF][START_REF] Rubas | Factors contributing to refrigerator cycling losses[END_REF] this phenomenon was estimated to result in an additional heat load of 2-7%.

Discussion

The thermographic technique

An infrared camera is a fast online method to analyze a temperature distribution. One of the advantages over conventional thermocouples technique is that it gives the full thermal landscape and not only point values. In the present study two sources of

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16 energy losses were found that would not easily have been found using thermocouples (if not, by chance, the thermocouples would have been located at the right spots). Some experiences from the experiments are that the test objects (evaporator and condenser) must be painted dull black to avoid thermal reflections. If possible, a wide angle lens should be used so that the evaporator could be filmed from a location within the cabinet. A thin plastic film is a useful tool to create a transparent "chimney" (see section 3.1). A limitation would be that only plate evaporators are suitable to analyze using IR-technique.

Identified losses

Four sources of energy losses were identified and will be discussed: 1. The evaporator activation process at startup. That is the time before the evaporator is fully charged with boiling refrigerant. 2. Evaporator edges poorly used for heat transfer during steady state. 3. Evaporation in filter dryer during the off-period. 4. Liquid refrigerant exiting the evaporator at startup.

1. The evaporator activation process at startup. Firstly, the evaporator studied was remarkably slow to activate at compressor startup (about 700 s). Even if this long time was partly a result of the experimental set-up the phenomenon is typical for household refrigerators. High evaporator superheat in the early on-period has been observed in earlier work [START_REF] Björk | Refrigerant mass charge distribution in a domestic refrigerator. Part I. Transient conditions[END_REF][START_REF] Coulter | An experimental analysis of cycling losses in domestic refrigerator-freezers[END_REF][START_REF] Krause | Cycling and quasi-steady behaviour of a refrigerator[END_REF]. Björk and Palm [START_REF] Björk | Refrigerant mass charge distribution in a domestic refrigerator. Part I. Transient conditions[END_REF] estimated this loss to be 3.6% for a household refrigerator similar to the one tested in the present study. A possible way to overcome this weakness and to speed up the activation is to redesign the evaporator refrigerant lines. Carrera [START_REF] Carrera | Influencing parameters on the refilling time of evaporators[END_REF] showed that evaporator lines in a downward inclination resulted in a much faster evaporator

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ACCEPTED MANUSCRIPT 17 activation. He also found the same result by decreasing the cross section area or to increase the mass flow by increasing the flow capacity in the expansion device. In another study by Björk and Palm [START_REF] Björk | Performance of a domestic refrigerator under influence of varied expansion device capacity, refrigerant charge and ambient temperature[END_REF] it was found that the energy consumption was independent of the expansion device flow capacity within certain ranges. Thus, it is suggested that a combination of increased expansion device flow capacity and rearranged refrigerant lines with a downward flow could be used to improve the energy efficiency. It should also be possible to reduce the internal refrigerant channel cross section area at the most upstream part of the evaporator without any significant additional pressure drop, as the vapor velocity at this location is low.

2. Evaporator edges poorly used for heat transfer during steady state.

Secondly, the IR-images show that the evaporator's edges were at higher temperatures than the central part of the evaporator. Thus, the edges were poorly used for heat transfer. To overcome this, the refrigerant channels should be moved closer towards the edges. Björk [START_REF] Björk | High efficiency cooling systems for refrigerators and freezers[END_REF] estimated the fin efficiency, calculated as the ratio between actual heat transfer and the heat transfer that would occur if the evaporator was isotherm at the refrigerant channel temperatures, of a similar evaporator to be 97%. This can be recalculated to a temperature loss of about 0.6 °C which corresponds to an efficiency loss of approximately 1.2% (1K added temperature difference is about 2% lowered efficiency for the given compressor).

3. Evaporation in filter dryer during the off-period.

Thirdly, the filter dryer acted as a liquid pocket during the off-period. In this pocket evaporation occurs resulting in a transport of thermal energy into the cooled space. It may be difficult to overcome this problem since the highly porous particles inside the M A N U S C R I P T A C C E P T E D ACCEPTED MANUSCRIPT 18 filter dryer are needed to absorb any water in the system. Assuming that the only water that enters the hermetically sealed cooling system comes during manufacturing, a more controlled manufacturing process would make the dryer (but not the filter) unnecessary. The loss was estimated to be 0.3 kJ/cycle or about 0.4% in efficiency (total cabinet heat leakage 73,6 kJ/cycle).

4. Liquid refrigerant exiting the evaporator at startup. Finally, the fourth source of energy loss was found at startup where it was observed that liquid refrigerant was exiting the evaporator. The classical solution to this problem would be to design the evaporator accumulator as a liquid separator so that boiling would solely occur in the evaporator. However, a liquid separator may not be easily designed within the dimensions of the plate. Moreover, if such separator was possible to design, it would still be unfavorable to boil refrigerant locally in the evaporator accumulator compared to have it boiling in the entire evaporator. The refrigerant needs to be distributed over the entire evaporator to give a high evaporation temperature and hence a high efficiency. Another solution would be to quickly recharge the evaporator after a startup. It has been found that a relation exist between the exiting refrigerant at startup and the evaporator activation time [START_REF] Björk | Performance of a domestic refrigerator under influence of varied expansion device capacity, refrigerant charge and ambient temperature[END_REF][START_REF] Rubas | Factors contributing to refrigerator cycling losses[END_REF].

With a large quantity of liquid exiting the evaporator at startup the liquid seal at the condenser outlet is quickly restored which permits faster evaporator activation. Moreover, a closer analysis suggests that the problem with liquid refrigerant exiting the evaporator to evaporate in the suction line heat exchanger (SLHX) at startup is not as large as it first appears. Part of this cooling is regenerated back to the cooled space via the SLHX itself. At similar conditions Björk and Palm [START_REF] Björk | Refrigerant mass charge distribution in a domestic refrigerator. Part I. Transient conditions[END_REF] calculated that the cooling energy accumulated in the SLHX material when exiting refrigerant evaporates

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19 at startup in the SLHX was about 2.0 kJ. This can be compared to 1.7 kJ of cooling energy accumulated in the evaporator material at startup, and the 1.7 kJ that was estimated to be the loss of cooling energy when liquid refrigerant liquid exits the SLHX at startup (passes un-evaporated through the SLHX). In other words, the cooling energy accumulated in the SLHX appears to be an important factor to decrease the cyclic losses. Perhaps a redesign of the SLHX, making the passage of liquid refrigerant at startup more difficult, would be a way to decrease the cyclic losses even more. Such design change should also be beneficial for the compressor that would receive less refrigerant liquid at startup. It is important, however, that the suction line pressure drop not increases with such design change.

Summary and Conclusions

Infrared thermography was used as an alternative technique to observe a household refrigerator cooling system during on-off cycling conditions. The experimental technique proved to be useful. It gave a clear overview of the cooling system temperature field. Overall, the experimental results confirm findings from previous research. It is for instance shown that refrigerant is displaced towards the compressor and condenser at startup which results in a starved, superheated evaporator and a drop in efficiency. However, the study also reveals sources to energy losses that would have been difficult to find if the conventional temperature measurement technique (i.e. However, it is limited to plate evaporators.

• The observed evaporator is slow to activate at startup. It is suggested to rearrange the refrigerant lines into a downward slope and to decrease the internal diameter at the upstream part of the evaporator.

• The observed evaporator edges are poorly used for heat transfer at steady state conditions. It is suggested that refrigerant lines are moved closer to the plate edges.

• Refrigerant evaporates in the filter dryer during the off period. In a more controlled manufacturing process the dryer could possibly be removed. However a filter is still needed.

• Liquid refrigerant exits evaporator at startup. However, it was calculated that a substantial part of this liquid is regenerated in the SLHX.

Acknowledgement

The Swedish National Energy Administration and Electrolux AB provided support for this work. This article is based on the experimental work by Nordenberg [START_REF] Nordenberg | Redistribution of migrated refrigerant[END_REF].

Animations

The IR stills can also be viewed as animations at the following web-addresses.

  cycling operation. The purpose is to assess an alternative technique for analyzing the transient behavior of household refrigerators.2.1 The typical refrigerator/freezerHousehold refrigerators and freezers typically operate by the vapor compression cycle.

Figure 1

 1 Figure 1 Images from the infrared videos captured five minutes after a compressor start. From the left: evaporator, the entire condenser and a close-up of the lower half of the condenser. The compressor is seen as a large warm object at the bottom of the condenser.

6 Cabinet: ( 1 .

 61 75 × 0.6 × 0.6 m), 0.350 m 3 internal volume, UA value 2.3 W/K (experimentally measured by a reversed heat leakage test). Evaporator: free convection, (0.66 × 0.49 × 0.0014 m), aluminium, plate type, back-wall located (20 mm distance to wall), integrated downstream located accumulator. Refrigerant line length (including accumulator) 6.22 m. Hydraulic diameter 3.2 mm. Total internal volume 116 ml whereof accumulator volume 46 ml in which approximately half the volume can store liquid at steady state condition. UA value estimated to 3.7 W/K. Weight 1.35 kg. Condenser: free convection, (1.33 × 0.51 × 0.008 m), steel, wire on tube (53 vertical wires on each side of the tubing, each of diameter 1.5 mm) located with 25 mm distance to the cabinet back wall. Refrigerant flow horizontally downward. Internal volume 135 ml. Internal/external diameter 3.5/5.0 mm. UA value estimated to 7.7 W/K Capillary tube: (2.54 m length and 0.6 mm internal diameter) with coaxial type suction line heat exchanger of 2 m length. The capillary tube adiabatic inlet and outlet sections are 0.5 and 0.04 m respectively.

Filter dryer:Figure 2 7 In

 27 Figure 2 Cooling system and thermocouple locations

Figure 3 IR

 3 Figure 3 IR-camera, AGEMA Thermovision 400 with a 40 mm/40° lens, in position for recording of the evaporator temperatures.

Figure 4 IR

 4 Figure 4 IR-camera in position for recording the condenser temperatures (close-up recording of condenser lower part)

  M A N U S C R I P T A C C E P T E D ACCEPTED MANUSCRIPT 9 and condenser images (calculated as root-sum-square of the individual uncertainties of the two films).

Figure 2

 2 Figure 2 Temperatures during one on-off cycle. For thermocouple locations see Figure 2.

  M A N U S C R I P T A C C E P T E D ACCEPTED MANUSCRIPT 10 evaporation pressure that causes the fixed volumetric compressor to operate with a high mass flow. The condenser temperature (T3) first rises sharply. It reaches a peak value of just above 40 °C in the middle of the on-period after which it slowly decreases to 40 °C just before the compressor stops. This decrease follows upon a steadily decreasing evaporation temperature which lowers the compressor capacity. The condenser outlet temperature (T4) parallels the condenser temperature (T3) but at a lower value. This is a clear sign of condenser sub-cooling.

  7 kJ/cycle corresponds to 4.7 g of refrigerant liquid passing un-evaporated through the SLHX into the suction line M A N U S C R I P T A C C E P T E D ACCEPTED MANUSCRIPT 13 just upstream of the compressor. This would explain the rapid cooling of the suction line observed.

•

  thermocouples) is used. It is shown that evaporator edges are poorly used for heat transfer and that evaporation occurs in the filter dryer during the off period. The identified losses are discussed and possible means to overcome them are suggested.It is concluded that: The experimental technique is useful to analyze the cooling system behavior.

First 12 Figure 1 Figure 2 Figure 3 IR-camera in position for evaporator filmingFigure 4 Figure 2

 1212342 Figure 1 Images from the infrared films five minutes after a compressor start. From the left: evaporator, the entire condenser and a close-up of the lower half of the condenser. The compressor is seen as a large warm object at the bottom of the condenser.

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


Yongliang Li

Haisheng Chen

Xinjing Zhang

Chunqing Tan

Yulong Ding
email: y.ding@leeds.ac.uk

Renewable energy carriers: hydrogen or liquid air / nitrogen?

Keywords: Energy carrier, energy storage, liquid air/nitrogen, hydrogen, renewable energy, ocean energy

The world's energy demand is met mainly by the fossil fuels today. The use of such fuels, however, causes serious environmental issues, including global warming, ozone layer depletion and acid rains. A sustainable solution to the issues is to replace the fossil fuels with renewable ones. Implementing such a solution, however, requires overcoming a number of technological barriers including low energy density, intermittent supply and mobility of the renewable energy sources. A potential approach to overcoming these barriers is to use an appropriate energy carrier, which can store, transport and distribute energy. The work to be reported in this paper aims to assess and compare a chemical energy carrier, hydrogen, with a physical energy carrier, liquid air/nitrogen, and discuss potential applications of the physical carrier. The ocean energy is used as an example of the renewable energy sources in the work. The assessment and comparison are carried out in terms of the overall efficiency, including production, storage/transportation and energy extraction. The environmental impact, waste heat recovery and safety issues are also considered. It is found that the physical energy carrier may be a better alternative to the chemical energy carrier under some circumstances, particularly when there are waste heat sources.

Introduction

Burning of fossil fuels as the main cause of some serious environmental issues such as greenhouse effect, ozone layer depletion and acid rains has long been recognized [START_REF] Veziroglu | 21st Century's energy: Hydrogen energy system[END_REF]. An obvious (and also hard) solution is to replace the fossil fuels with renewable ones such as wind, solar, ocean energy etc [START_REF] Alnaser | Estimation of energy from tide, wave and sea water currents in Bahrain[END_REF]. Currently, the renewable energy resources account for ~ 5% of global power capacity and 3.4% of global power generation excluding large hydropower (which is about 15 % of the global power generation) [START_REF]Renewables[END_REF]. Like a number of other countries, the UK government has set a target for electricity generation from renewable resources to be increased from about 4.6% at present to 20% by 2020, and EU proposed recently an even higher goal of between 30 and 40% [START_REF] Government | Meeting the Energy Challenge: a white paper on energy[END_REF].,As a consequence, significant efforts have been made in the renewable energy research and these efforts have been mostly focused on solar, wind, biomass and geothermal sources. While these sources are very promising, the world's oceans could emerge as an important source to provide economically viable renewable sources [START_REF] Pelc | Renewable energy from the ocean[END_REF].

Ocean winds blow harder and with more reliable consistency than the wind on land, which more than offsets the greater cost of building offshore wind power facilities. Oceans tides also contribute massive amounts of renewable energy that is gravitationally derived through the interplay of the earth and the moon. The energy from ocean waves and tidal streams, along with ocean-based wind energy, make the world's oceans a source of renewable energy that may in the next few decades greatly outstrip solar energy as an economical alternative.

However, renewable energy sources particularly wind and solar are intermittent and they often do not match the actual energy demand. This makes the energy storage and power management increasingly important. There have been reports that the amount of wind energy that could be utilized is seven times greater than the current situation if a stable storage facility is available [START_REF] Merriam | Wind, waves and tides[END_REF]. Energy storage refers to a process for storing energy produced at times of low demand and low generation cost and releasing at times of high demand and high generation cost or when there is no more generation capacity available. The stored energy can be in the form of either potential energy (e.g. chemical, gravitational or electrical energy) or kinetic energy (e.g. thermal energy or motion). This paper is concerned about the use of ocean renewable energy resources. A potential issue for the use of the ocean resource is the remote locations [START_REF] Meisen | Linking Renewable Energy Resources Around the World: A Compelling Global Strategy[END_REF], which requires high capital, operation and maintenance costs if the conventional electricity transmission and distribution approaches are used. An alternative is to use an energy storage medium, which can be transported by ships. For such an approach to work the energy storage medium is the key. An ideal energy storage medium is the one that has a high energy density and can be detached easily and completely from the generation and release

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3 devices so that they can be transportable. In this sense, the energy storage medium can be regarded as an energy carrier, and the conventional methods of large scale energy storage approaches are less useful. For example, the pumped hydroelectric storage (PHS) method stores the gravitational energy of water, which is not suitable for transportation. The compressed air energy storage (CAES) method stores energy in the form of high pressure air, which has a low energy density and is hence too bulky to transport.

It is well known that hydrogen has been regarded as a popular carrier of renewable energy in remote locations [8; 9; 10].

Using such a chemical energy carrier, the ocean energy generates electricity first, which is then used to electrolyze water to produce hydrogen. Hydrogen is then transported to the end-users for transforming back to electricity or kinetic energy by a fuel cell or other devices. Cryogens normally refer to as liquid media at a temperature below ~ -150℃ and are physical energy carriers [START_REF] Ahern | Applications of the 2nd Law of Thermodynamics to Cryogenics -a Review[END_REF]. Examples of the cryogens include liquefied natural gas (LNG), liquid air and liquid nitrogen. The use of liquid air/nitrogen as an energy carrier is not commonly known until fairly recently [12; 13; 14; 15]. Energy storage using liquid nitrogen / liquid air involves the production of the cryogen through an air separation and liquefaction process. The cryogen is then transported to end-users where it is heated using the environment heat, waste process heat or heat from renewable resources (e.g. solar) if available and expands to generate electricity using a cryogenic heat engine. As a renewable energy carrier, cryogen can also provide direct cooling and refrigeration, air conditioning units and acts as power source for vehicles and ships.

Fig. 1 illustrates schematically the use of ocean energy with hydrogen and cryogen as the energy carriers. This paper assesses and compares the performance of two energy carriers mainly for the use of ocean renewable energy though the results are expected to be applicable to other renewable energy resources. Liquid air and nitrogen will be used as the model cryogens. The assessment includes production, storage/transportation and energy extraction processes as well as their environmental impacts and safety issues. Note that hydrogen in its liquid form is also a cryogen. For clarity and conciseness, liquid hydrogen will not be referred as a cryogen in this paper.

Carrier production

Two main technologies used to produce hydrogen are considered in this work, reforming reaction and water splitting. The most common hydrogen production method in commercial use today is the former, which uses hydrocarbons or other chemical compounds such as coal and biomass as feed stocks. However, as the renewable resources considered here (wind,

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4 tides and waves) contain only the mechanical energy of moving masses of air or water, direct chemical path of hydrogen production is to split water using electrical energy generated by the kinetic energy of renewable resources. Therefore, hydrogen production using the ocean resources can be described as follows. The kinetic energy of moving masses (air/seawater) is extracted first by mechanical machines to form the mechanical energy of the machines. The fraction of extractable power depends on the form of energy and extraction processes and devices, and is limited by a theoretical value determined by the thermodynamics. For example the upper bound of the wind energy extraction by an ideal horizontal axis machine is 0.593 (power coefficient) under some rather general assumptions. Then the mechanical energy of the mechanical machines is converted to electricity with a conversion efficiency of about 75-95%. Finally, the electrical energy is employed to produce hydrogen by electrolysis. Table 1 summarizes the performance of electrolysis technologies, along with their feedstock and efficiencies (defined as the low heating value of hydrogen produced divided by electrical energy consumed in the electrolysis cell) [16; 17; 18]. In the summary table, the high temperature electrolysis efficiency is dependent on the temperature at which the electrolyzer operates and the efficiency of the thermal energy source. If everything is considered, the efficiency of converting mechanical energy to chemical energy (hydrogen) is within a range of 37.5 -66.5%.

As mentioned above, the cryogen production is done by the air separation/liquefaction process in which cryogenic or Stirling engine coolers liquefy the main components of air through the well-known Joule-Thomson effect, see Fig. 2. This is different from the hydrogen production process as discussed above as both the compression and refrigeration processes in the liquefaction process could be directly powered by the mechanical work of the ocean energy. Therefore cryogen production could be more competitive than hydrogen production in terms of the overall efficiency and capital costs as the electricity conversion process is essential in the cryogen production end. However cryogen production is an energy-intensive process. Currently the efficiency of practical air liquefaction plants strongly depends on the plant scales; see Table 2 for the typical data of cryogenic industry [START_REF] Chen | An Overview of Air Liquefaction and Separation Process[END_REF]. Note that the efficiency of air separation / liquefaction plants with a capacity of few tons/day of liquid nitrogen are about 50% [START_REF] Franz | Cryogenic Heat Engines Made Using Electrocaloric Capacitors[END_REF], and there are reports that the efficiency could be further enhanced to 60% and higher [21; 22; 23].

Carrier storage/transportation

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As the energy carriers are produced at the offshore power generation facility, they have to be stored and transported to an onshore location for distribution to the end-users. Hydrogen could be delivered to onshore facilities through one of the three forms: gas (compressed), liquid (liquefied), or solid (in a solid hydrogen carrier). In the gaseous form, hydrogen can be compressed and transported to onshore in pressurized containers. Because of its low molecular weight, hydrogen molecules are very small and leakage can be an issue particularly at high pressures. Because of this, the storage pressure of hydrogen is limited to ~ 35 MPa at present. Liquefaction of hydrogen and its transportation in containers are an established technology.

However, the liquefaction process is energy-intensive and about 30-40% of the energy content is lost in the liquefaction process [START_REF] Usde | National Hydrogen Energy Roadmap[END_REF]. Because of its added complexity and costs for both generation and transportation, the use of liquid hydrogen is not regarded as an attractive option.

Table 3 shows the volumetric capacity of hydrogen under different conditions. The energy density increases linearly with increasing storage pressure, and liquid hydrogen has a higher value. As mentioned above, high-pressure storage of hydrogen gas is limited by the weight of the storage canisters and possible leakage development. This imposes potential safety issues [START_REF] Reider | Hydrogen safety problems[END_REF]. A promising way to replace the conventional hydrogen fuel storage methods is to use a solid state hydrogen carrier, which refers to any substance that can store and transport hydrogen in either a chemical or a physical state.

The advantages of using solid carriers are better safety and reliability in comparison with liquid or compressed gas storage methods. The carrier is charged with hydrogen at the offshore generation site and transported to onshore where hydrogen is stripped off. The carrier needs to be taken back for recharging (two-way carrier) or decomposed at the point of hydrogen use (one-way carrier). Examples of hydrogen carriers include ammonia (one-way) and liquid hydrocarbons and metal hydrides (two-way). It is reported that some of the hydrogen carriers could attain a volumetric capacity as high as liquid hydrogen [26; 27]. However the weight of current hydride substrates and their container is much greater than that of the stored hydrogen, so extra energy is required during both charging and discharging processes. While considerable research has been done on the hydrogen carrier technology, no reports have been found so far on the commercial use of the carriers.

Compared to hydrogen, the delivery of cryogen is much easier. Once produced and stored in insulated containers, cryogen is ready to be delivered. No extra energy required except for the pumping power consumption which is negligible for a liquid. As shown in Table 3, the volumetric energy density of cryogen is much lower than liquid hydrogen but at the same order of magnitude with compressed hydrogen. It should be noted that the volume-based energy densities of both cryogen and compressed hydrogen gas (at practicable pressures) are significantly lower than that of traditional fossil fuels.

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The only energy loss of cryogen is the heat dissipation of cryogenic tank at the ambient pressure which can be less than 1% per day in an insulated Dewar using conventional insulation technologies.

Energy extraction

Upon delivered to end users' side, the energy stored in the energy carriers is extracted. Current technologies converting the chemical energy of hydrogen to mechanical energy or electricity energy use one of the two methods, combustion and electrochemical conversion in e.g. a fuel cell.

For the combustion method, both hydrogen internal combustion engines and hydrogen fuelled gas turbines have been investigated [28; 29; 30; 31]. Because of high burning temperature, hydrogen internal combustion in a conventional engine produces a very high level of nitrogen oxides which cause environmental problems. Although there are a number of ways to decrease nitrogen oxides emissions, this will decrease the efficiency of the engine at the same time. This is serious as currently the low heating value (LHV) efficiency of internal combustion is only 20~35%. Another approach is the use of gas turbines in hydrogen fueled combustion prior to which pure oxygen is generated from an air separation unit to avoid the production of nitrogen oxides. In such a process, liquid water is often used as a mixing fluid to decrease the turbine inlet temperature, see Fig. 3 (a). Here water is preheated by the exhaust gas from the gas turbine. Different from the internal combustion, the burning temperature of these cycles is limited by the turbine inlet temperature (TIT) which is currently about 1300℃. In these cycles the exergy loss mainly is caused by preheating and mixing of the three input stream reactants in the combustor and accounts for about 40~50%. This, plus the energy consumed by pure oxygen production, the overall LHV efficiency of this type of cycles is limited to about 50% using current technologies and may increase to about 60% if the turbine inlet temperature is increased to ~2000℃ with future developments in turbine and material technologies [START_REF] Jin | A novel gas turbine cycle with hydrogen-fueled chemical-looping combustion[END_REF].

A different approach to the direct combustion of hydrogen is the so-called chemical-looping method for extraction of hydrogen chemical energy as shown in Fig. 3 (b). Such a method uses two successive reactions, metal oxide reduction with hydrogen, and subsequent oxidation of the metal by air, yielding the metal oxide and a high-temperature flue gas. The resulting high temperature gas is then used to power turbines. By adding a chemical-looping process, the exergy loss in the combustion process decreases significantly [33; 34]. It is claimed that the LHV efficiency of hydrogen in such a cycle could be as high as 63% [32; 35].

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Another approach for enhancing hydrogen energy extraction efficiency is the use of fuel cells, which could replace internal combustion engines and turbines as a primary way to convert chemical energy to kinetic or electrical energy. Fuel cells work via electrochemical principles, and hence are more efficient than heat engines. Table 4 [36; 37; 38] lists the technical information of three types of hydrogen fuelled fuel cells including typical efficiencies, operating temperatures, catalysts and other operating parameters, where the efficiency refers the cell efficiency and the system efficiency is generally 10% lower. However, there are still a number of barriers for the industrial take-up of the fuel cell technologies; these include requirements for high purity of hydrogen, high costs to manufacturing, low cell reliability and short service life. For example, the cost of fuel cell technologies is about 4-5 times more expensive than the combustion engines / turbines, while its service life 2-3 times shorter [START_REF] Bennaceur | Hydrogen: A future energy carrier? Oilfield Review[END_REF].

In contrast to hydrogen, energy extraction from cryogen is much more straightforward and easier. Liquid air/nitrogen is heated and expands to produce power. This is similar to a steam engine but expansion of cryogen does not emit pollutants such as NO X and particulate matters etc. Heating of the cryogen uses the free environmental heat. However, the exergy efficiency of this approach is limited to about 40% as a considerable portion of the exergy is in the form of cold, which is wasted during the heating process. In order to improve the efficiency, combined cycles are studied to recover the waste cold energy by integrating the direct expansion of cryogen with other cycles, see Fig. 4 for an example where a direct expansion-Rankine hybrid cycle with propane as the working fluid for the Rankine cycle. The combined methods for cryogenic energy extraction has been discussed in detail by the authors in Ref [START_REF] Li | Fundamentals and applications of cryogen as a thermal energy carrier: A critical assessment[END_REF]. It has been concluded that the combined cycle cryogenic heat engines could have an exergy efficiency of 65~78% under ideal conditions [12; 41].

The exergy efficiency of cryogen expansion could be further enhanced if the working fluid is superheated by waste heat or heat from other renewable sources such as solar. At present, only high grade waste heat in gas turbine power cycles could be recovered efficiently by heat recovery steam generator (HRSG) technology while low grade heat is generally vented. The use of a cryogen expansion cycle can be a very effective way to recover such low grade heat as the boiling point of air or nitrogen is much lower than the ambient temperature. Chen et al. [START_REF] Chen | A method of storing energy and a cryogenic energy storage system[END_REF] claimed that the exergy could be improved from ~78% to 117% if the working fluid were superheated to ~100 o C, and the exergy efficiency could be doubled if the waste heat is as high as 300 o C. Note that here the waste heat is regarded as a free energy source in the calculations by Chen et al. [START_REF] Chen | A method of storing energy and a cryogenic energy storage system[END_REF]. However, cryogen may be more competitive than hydrogen as an energy carrier in terms of capital costs because there is no electricity conversion at the cryogen production end.

Concluding remarks and outlook

• As an energy-intensive process which consumes up to ~ 40% of the chemical energy, pretreatment is required for the storage and transportation of hydrogen regardless of the form of the carrier being compressed gas, liquefied hydrogen or solid hydrogen carrier. In addition, the hydrogen carrier transportation may be two-way if the two-way carrier is adopted. In contrast, transportation of cryogen requires neither pretreatment nor other carriers but insulated containers. However it should be noted that hydrogen has a higher volumetric energy density than cryogen regardless of storage forms.

• The energy extraction efficiency of hydrogen depends significantly on the conversion methods. Currently, gas turbine and fuel cell methods have better performances than conventional internal combustion engines though significant developments are needed in order for the fuel cell technologies to be competitive. Cryogen engines, if cold recovery technology is employed, may have an even higher efficiency than the fuel cell technologies.

• Waste heat especially low grade heat could be recovered efficiently by the use of cryogen engines.

• Although hydrogen is regarded as a clean fuel, nitrogen oxides can be produced if air is used as an oxidant.

Cryogen is much more environmentally friendly as the processes of production (using the ocean energy) and extraction are both physical involving no chemical transformation. Even the transportation of the cryogen can be cryogen-based.

Overall, hydrogen is viable as an ocean energy carrier only if a practical method to store and carry hydrogen is introduced and fuel cells become cheap, more tolerant to impurities of hydrogen, and with their service life much improved. Before these are achieved, cryogen appears to be a more attractive energy carrier as there are few technical difficulties to overcome. Moreover, the overall efficiency of the use of cryogen can be greatly increased if low grade heat is used in the process of cryogen energy extraction. The low grade heat can be obtained from either traditional heavy industrials like power plants or directly from the abundant renewable solar energy. As mentioned before, the electricity conversion efficiency of solar energy is limited; it is much easier to use solar thermal energy in countries blessed with good sunshine (e.g. India with a mean daily solar radiation of 5-7kW/m 2 [START_REF] Sharma | Solar cooker with latent heat storage systems: A review[END_REF]) via e.g. a cascading way. Solar energy, along with the thermal M A N U S C R I P T A C C E P T E D ACCEPTED MANUSCRIPT 9 energy storage technology using e.g. phase change materials, has been extensively studied and used for heating, cooking and even electricity production [42; 43; 44; 45]. Therefore, a solar-cryogen hybrid energy storage system appears to be a promising way for both transportation system and electricity production. In such a system, solar energy is stored in the form of heat in e.g. phase change materials and other renewable energy like ocean energy is stored in cryogen. After transported to end users, the two energy sources are combined to produce kinetic power or electricity, using the cryogen as the working fluid. As both the heat and cold sources are generated by renewable energy sources and there are no chemical reactions involved, the system can be very environmental friendly and sustainable.

  and compares two energy carriers, hydrogen and cryogen, for exploitation of ocean energy sources. The assessment and comparison are based on the overall efficiency including production, storage/transportation and energy extraction. The environmental impact, waste heat recovery and safety issues are also touched upon. The following observations are drawn:• The production efficiencies of hydrogen and cryogen are similar at 40 ~ 65% based on the current technologies.

Acknowledgements

The authors would like to thank the following agents for financially supporting the work: (i) UK EPSRC for a Dorothy Hodgkin Award (YL and YD), (ii) UK EPSRC for a grant under EP/F060955/1 (HC and YD), (iii) Institute of Process Engineering of Chinese Academy of Sciences under a collaborative energy research initiative (YD), and (iv) China Scholarship Council for a scholarship to Institute of Engineering Thermophysics (XZ and CT)

Table 1

Efficiencies of the different electrolysis technologies

Table 2

Real world efficiency for liquid air/nitrogen production in typical cryogenic industry


Jérôme Barrau

Daniel Chemisana
email: daniel.chemisana@macs.udl.cat

Joan Rosell
email: rosell@macs.udl.cat

Lounes Tadrist
email: lounes.tadrist@polytech.univ-mrs.fr

M Ibañez
email: m.ibanez@macs.udl.cat

An experimental study of a new hybrid jet impingement / micro-channel cooling

Keywords: ATE Jet impingement, Microchannel, Heat sink, Temperature uniformity Nomenclature

A new hybrid cooling scheme is proposed for high heat flux management and power devices. This scheme combines the benefits of micro-channel and jet impingement cooling technologies, with the additional objective of improving the temperature uniformity of the cooled object. The geometry is tested experimentally to characterize its performances and to assess its capacity to provide this characteristic. The temperature distribution of the heat sink has been measured through a matrix of thermocouples. The hybrid cooling scheme is shown to have the capacity to optimize the temperature uniformity of the cooled object, since the experimental design provides a global decrease of the temperature of the heat sink in the direction of the fluid flow.

At area of aluminium block's top test surface (m 2 ) Cp specific heat (J/kg•K) d h hydraulic diameter (m) ∆P in-out pressure losses between the inlet and the outlet of the heat sink (Pa) ∆T in-out temperature difference of the fluid between the inlet and the outlet of the heat sink (K) h l local heat transfer coefficient (W•m 2 /K) k thermal conductivity (W/m.K) L length (m) R thermal resistance coefficient (K•m 2 /W) T temperature (K) T S temperature of the surface to cool (K) T F temperature of the fluid in the exchange interface (K) T in inlet temperature of the refrigerant fluid (K) T out outlet temperature of the refrigerant fluid (K) T W average temperature on the wall in which is applied the heat flux (K) P f power absorbed by the refrigerating fluid (W) P W power delivered by the feeding source (W) q" heat flux (W/m 2 ) q" eff effective heat flux based on top test surface area of aluminium block, q" eff =P i /A t (W/m 2 ) Q flow rate (m 3 Several studies have been published on the fluid flow and heat transfer characteristics of single-phase high heat flux removal schemes. These cooling schemes, which use micro-channel and jet impingement systems, are commonly employed in the cooling of high-performance microprocessors, laser diode arrays, radars, X-ray anodes, and solar photovoltaic concentrators. Microchannel and jet impingement schemes have recently been compared [START_REF] Kandlikar | Evaluation of jet impingement, spray and microchannel chip cooling options for high heat flux removal[END_REF][START_REF] Lee | Comparative analysis of jet impingement and microchannel cooling for high heat flux applications[END_REF] and are the object of constant developments. Both technologies provide very high flux removal, but still present serious drawbacks. Micro-channels cause a large increase in temperature along the path of the fluid flow and are also responsible for a relatively large pressure drop. Jet impingement requires the use of numerous jets to ensure a relatively good level of uniformity of temperature across the base of the heat sink. These cooling schemes have the disadvantage of reducing the local heat transfer coefficient because of the interaction between adjacent jets. The solution to this problem lies in the use of complex geometries that allow the extraction of the spent flow from the jets. Many studies have attempted to reduce these effects [START_REF] Vafai | Analysis of two-layered micro-channel heat sink concept in electronic cooling[END_REF][START_REF] Chong | Optimisation of single and double layer counter flow micro channel heat sinks[END_REF][START_REF] Royne | Design of a jet impingement cooling device for densely packed PV cells under high concentration[END_REF] and at the same time to enhance the heat transfers associated with the two technologies.

Temperature non uniformities in micro-channels or jet impingement cooling affect the performance of electrical systems and causes mechanical stresses due to the difference of thermal expansion from one side to the other of the cooled object. Both factors reduce the reliability of the whole system. As a consequence, main research on heat sinks focuses on improving the resistance of the devices to thermal cycles, matching the thermal expansion coefficient of the "packaging" to the one of semiconductors. Another way of investigation is to improve the temperature uniformity of the cooled object.

To date, few works have been published about hybrid systems in which the benefits of micro-channel and jet impingement are combined to overcome the main drawbacks associated with the two techniques. Jang et al. [

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experimentally investigated a micro-channel heat sink subjected to an air impinging jet. They demonstrated that, under certain specific conditions, the efficiency of cooling was 48.5% better than that achieved applying a classical micro-channel design, with pressure losses being reduced by 90.5 %. The manifold micro-channels developed by Copeland et al. [START_REF] Copeland | Manifold microchannel heat sinks: Theory and experiment[END_REF] and optimized by Ryu et al. [START_REF] Ryu | Three-dimensional numerical optimization of a manifold microchannel heat sink[END_REF] is another technology that has a similar effect to that of impinging jets on a micro-channel heat sink.

The combination of both technologies in hybrid cooling schemes makes it possible to significantly increase thermal exchange within the cooling system. In effect, this design allows to reduce the length of the flow circuit by the half in the dissipation system. This is important, as indicated by Lee and Vafai [START_REF] Lee | Comparative analysis of jet impingement and microchannel cooling for high heat flux applications[END_REF], since an increase in the length of the flow circuit causes a reduction in the maximum quantity of heat that can be extracted by the system, whether through microchannels or through jet impingement devices. Furthermore, this shortening of the flow circuit implies a reduction in pressure losses within the system [START_REF] Kandlikar | Extending the heat flux limit with enhanced microchannels in direct single-phase cooling of computer chips[END_REF]. An other important feature of jet impingement / micro-channel hybrid systems studied in the literature is that they were shown to improve the uniformity of the temperature of the cooled surface. Sung and Mudawar [START_REF] Sung | Experimental and numerical investigation of singlephase heat transfer using a hybrid jet-impingement/micro-channel cooling scheme[END_REF][START_REF] Sung | Single-phase and two-phase cooling using hybrid microchannel/slot-jet module[END_REF] analyzed hybrid jet impingement / micro-channel cooling schemes both experimentally and numerically (Fig. 1) and conclude that the hybrid modules maintain a higher degree of surface temperature uniformity than other cooling scheme. Mudawar [START_REF] Sung | Experimental and numerical investigation of singlephase heat transfer using a hybrid jet-impingement/micro-channel cooling scheme[END_REF] In this study, a new hybrid cooling scheme is experimentally evaluated. This scheme involves feeding coolant from a slot jet in a modified micro-channel heat sink structure. Experimental analysis is used to study the main trends of the heat exchanges of the hybrid cooling scheme and to assess the capacity of the heat sink to improve a high temperature uniformity of the cooled object.

2 Controlled temperature hybrid jet impingement microchannel cooling scheme

Conceptual analysis

In all cooling schemes, the temperature of the refrigerating liquid (T f ) increases throughout the circuit as a result of the absorption of energy. In the case of micro-channels, the heat transfer coefficient remains constant throughout the whole heat sink due to the geometry of the system while impinging jets offer a characteristic global reduction in the heat transfer coefficient along the circuit. However, as the energetic flux (q") that arrives at the cooling system is generally uniform, the temperature at the base of the heat sink (T S , l ) increases in the direction of the fluid flow.

The local heat transfer coefficient (h l ) is calculated from equation (1):

l f l S l l T T q h , , " - = (1)
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The local temperature of the bottom of the heat sink can be extracted from equation (1):

l f l l l S T h q T , , " + = (2)
Analyzing this equation, it is evident that increasing the heat transfer coefficient throughout the circuit is the only way to obtain a uniform temperatures pattern on the heat sink's base (T S ).

In both micro-channels, and impinging jets, thermal exchanges vary according to the regime of the flow and the geometry of the circuit. For a given design, the flow regime remains fixed by the fluid velocity and the inlet section of the heat sink. Therefore, the only way to increase the heat transfer coefficient is to develop a cooling scheme with a variable internal geometry in the direction of the coolant flow.

Design description

On the base of the bibliographic analysis of the different parameters of cooling systems involving micro-channels and impinging jets, a new design is proposed (Fig. 2).

Fig. 2. Proposed design (linear symmetry)

In this scheme, fluid enters through the slot located in the symmetry plane of the heat sink (arrows in blue), between the two upper plates (which were cut in order to enable us to observe the geometry of the system) and leaves it through the extremities of the channels (arrows in red). Between entering the heat sink and subsequently exiting it, the fluid flows through a series of microchannels with a variable longitudinal distribution. The linear symmetry of the system means that both the entry and exit of the liquid are simpler than under conditions of circular symmetry. Furthermore, this guarantees the symmetry of thermal exchange in one dimension. This design entails two well differentiated areas. The first corresponds to a cooling system using jet impingement, while the second relates to a system of micro-channels. The relative arrangement of the channels with regard to the impact area of the fluid defines the degree of interaction between the two types of thermal exchange.

Besides combining different characteristics that directly influence local heat transfer coefficients, this system constitutes a change from the matrices of impinging jets and also from the hybrid jet impingement / micro-channel system suggested by Sung and Mudawar [START_REF] Sung | Experimental and numerical investigation of singlephase heat transfer using a hybrid jet-impingement/micro-channel cooling scheme[END_REF]: In effect, in both types of cooling system, the flows from adjacent jets meet head on, implying a major deceleration of flow and even points where fluid flow stops and becomes stationary. In the proposed design, the radial acceleration of the fluid that impacts against the surface to cool, which is an essential aspect of the physics of thermal exchange at jet impingement heat sinks, is free.

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3 Experimental Set up

Experimental device

The relatively high thermal conductivity of aluminium (k Al =237 W/m.K at 300 K) and the simplicity of its mechanization led us to make the experimental device using this material (Fig. 5). The upper platform of the experimental heat sink, which corresponds to the area of the system to be cooled, measures 2.55 cm (width) by 5.95 cm (length). The depth of the micro-channel, which are fitted at the top of the heat sink, is 2.5 mm. Their widths vary at different points along the refrigerating liquid circuit, ranging from 3.5 mm to 0.5 mm, with an intermediate section in which the channel width is 1.5 mm. The fins, which serve as vertical walls in the micro-channels, are 0.5 mm wide.

In the lower part of the system, a housing to every part end was added, in order to introduce electrical resistances that simulate energy that must be extracted by the heat sink. These housings are separated from the bottom of our dissipation system by a thickness of 23 mm that made it possible to homogenise the thermal flow within the mass of material and thereby ensure that it arrives at the point of thermal exchange with the refrigerating liquid in as uniform a state as possible. As shown in Fig. 3, the 18 1.1 mm diameter drill holes observed in the front view of the device, under the bottom of the heat sink, are used to house the 1 mm diameter type K thermocouples (Chromel-Alumel). The depth of these drill holes is half the width of the experimental heat sink in order to measure the temperature of the system along the longitudinal plane of symmetry of the device.

At the top of the experimental heat sink, the refrigerating liquid distributor in our cooling system is formed by a single Plexiglas part composed of three volumes: -One entry volume, which guides the refrigerating liquid to the area of the jet impingement of the heat sink through a slot in its bottom. -Two exit volumes, which collect the liquid leaving the micro-channels at the extremities of the heat sink.

The lower face of this part, which is in contact with the fins of the heat sink, is used to create the closed micro-channels through which the refrigerating liquid flows. Drill holes were made in each of the three volumes of the space used for distributing the refrigerating liquid in order to allow the introduction of pressure and temperature sensors. Finally, a Plexiglas cover was fitted to the top of the distributor hall with a drill and tightened onto each of the three volumes in order to receive the connectors of the hydraulic circuit.

Experimental mount

The experimental device was then secured to a desk using a support made with an insulating material (Bakelite). The rest of the experimental model was insulated using rock wool.

The hydraulic circuit is designed to provide temperature, pressure and flow control at the point at which the refrigerating liquid enters in the system (Fig. 4). The 3 pressure sensors are located in both exit volumes and in the entry one.

Fig. 4. Experimental set up

In order to obtain temperatures from the bottom of the dissipation system from several sections along the fluid circuit (T zo,j with j=0,1,2,3,4,5) without intrusive measurements that disturb the flow and consequently invalidate any results obtained, the methodology described by Robinson and Schnitzler [START_REF] Robinson | An experimental investigation of free and submerged miniature liquid jet array impingement heat transfer[END_REF] was used. Vertical series of thermocouples with known vertical and transversal positions were located in every section where the measurement of the temperature at the bottom of the heat sink was planed (Fig. 5). This matrix of thermocouples (T ij , with i=1, 2, 3 and j=0, 1, 2, 3, 4, 5) allows determining the temperature at the bottom of each section through the following calculation:

5 , 4 , 3 , 2 , 1 , 0 and 3 2 for
) ( 1 1 1 0 1 , 0      = = =         - - ⋅ - + = j or i i z z T T z z T T j ij j ij j j j j z
(3)

Fig. 5. Controlled parameters of the experimental module

The uncertainties in flow rate measurements, pressure levels, electrical power levels and temperatures were ±2%, ±1%, ±10 W and ±0.2 ºC, respectively.

Results and discussion

Energy Balance

A sequence of measurements was considered valid if the energy balance of the system was verified: all the power delivered by the feeding source (P i ) and therefore transferred, in the form of heat, to the experimental model through the resistances, must have been absorbed by the refrigerating liquid (P f ). This power is calculated through the expression:

( )

in out water water f T T Q Cp P - ⋅ ⋅ ⋅ = ρ (4)
At the end of the calculations, the energy balance is verified if:

W f P P = (5)
The results of the energy balance can be presented in terms of the observed difference in the temperature of the refrigerating liquid between the inlet and the outlet of the heat sink (T out -T in ) and the Reynolds number. The

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section through which the fluid enters the heat sink is an elongated slot with a form factor ξ defined by the expression:

jet jet L W = ξ (6)
In such sections, the hydraulic diameter is calculated according to:

ξ + ⋅ = 1 2 d h jet W (7)
The Reynolds number is calculated using the expression:

υ h d . v Re = (8)
The conditions for the seven validated tests are presented in table 1.

Table 1

Tests conditions Fig. 6 compares the measured water temperature rise between the channel inlet and outlet, and theoretical values predicted from the simple energy balance. The excellent agreement between the two proves virtually all the electrical power supplied by the cartridge heaters was removed by the water, and heat losses are indeed negligible.

Fig. 6. Comparison of experimental data and energy balance predictions

for water temperature rise from heat sink inlet to outlet

Temperature maps

The qualitative observation of the evolution of temperature at the bottom of the heat sink, for the different test conditions, indicates that the studied cooling scheme solves one of the features initially suggested as a design criterion (Fig. 7). Indeed, the temperature at the bottom of the heat sink shows an overall downward slope throughout the fluid circuit, indicating that by optimizing the parameters of the design, it is possible to achieve a uniform temperature throughout the heat sink. The average temperature at the bottom of the heat sink (T z0,m ) is calculated by the equation:

∑ = = ⋅ + = j j z j z m z T T T ( 9
)
where T z0,j , for j = 0, 1, 2 , 3 , 4 ,5 are the temperatures at the bottom of the heat sink in the experimental module sections S 0 , S 1 , S 2 , S 3 , S 4 y S 5 respectively. Table 2 presents results for temperature uniformity at the bottom of the heat sink. They are expressed by the standard deviation (σ T ) with respect to average temperature and by the maximum temperature difference on this surface.

Table 2

Temperature uniformity at the bottom of the heat sink The temperature uniformity does not decrease linearly with regard to differences in the temperature of the fluid between the inlet and the outlet of the heat sink (Fig. 8). This means that, although the fluid warms up during its path through the circuit, the system maintains a thermal flow extraction with a similar uniformity. This is a good indicator of the optimization of the design with regard to the criterion of temperature uniformity.

Heat exchange

Calculation of the average thermal resistance coefficient m t R of the system was carried out by applying the following equation:

eff in w m t q T T R " ) ( - = (10)
where q" eff is the heat flux , T in the water temperature at entry to the heat sink and T w the average temperature of the wall on which the heat flux took place. In this study, in order to simulate the thickness of the object to be cooled, this average temperature was calculated at the first row of thermocouples (z = z 1 = -1.6 mm). The average thermal resistance coefficient for the four different tests conditions is presented in table 3.

Table 3

Average thermal resistance coefficient and pressure losses Fig. 9 shows the distribution of the local Nusselt number along the heat sink bottom wall. The local Nusselt number is defined as

) ( " in S f h x T T k d q Nu - ⋅ ⋅ = ( 11
)
where q" is the local heat flux obtained by the equation

1 3 ) ( " 1 3 z z T T k q z z Al - - ⋅ - = . ( 12
)
Fig. 9. Nusselt number distribution on the heat sink bottom wall. Error propagation is included in the calculation of the Nusselt number (1.36%).

The effect of jet velocity in the jet impingement influence zone is well shown with the Nusselt number distribution. Indeed, for high flow rates (test 1 and 2),

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the Nusselt number remains practically constant from x = 0 to 11 mm. Furthermore, it's very interesting to note that, if the local Nusselt number is multiplicated by the number of channels in each section, the results will become in a practically constant trend from x = 0 mm to x = 27,5 mm. This characteristic implies the quite uniform temperature profile at the heat sink bottom wall (Fig. 7). The local Nusselt number of the proposed hybrid jet impingement/microchannel cooling scheme (with water as coolant fluid) is higher than the one of the hybrid micro-channel / jet impingement module calculated by Sung and Mudawar [START_REF] Sung | Single-phase and two-phase cooling using hybrid microchannel/slot-jet module[END_REF]. Specifically, for Re=2892, q" eff =20.58 W/cm 2 with HFE-7100 as working fluid, the maximum local Nusselt number is 56.19, with higher longitudinal (x) and transversal (y) non uniformities of the local Nusselt number distribution. In effect, the temperature uniformity of the proposed design in the y axis is assured by the geometry of the heat sink.

Pressure losses

Fig. 10 shows the measured pressure losses of the system for the seven tests presented. The quantitative values are lower than for micro channels devices in the same configuration. In fact, the large pressure losses of the micro-channels are due to the large contact area between the fluid and the walls of the heat sink. This implies great friction. In the proposed hybrid system, the contact area between the fluid and the walls only appears in the final part of the circuit, which implies smaller pressure losses. The increase of the pressure losses with the Reynolds number in a non-linear form was observed, according to trends derived from Moddy's diagram.

Conclusion

After the analysis of the bibliography related to hybrid jet impingement / micro-channels heat sinks, a new cooling scheme has been proposed. The single-phase performance was experimentally studied in pursuit of a geometry that yields lower surface temperature and smaller temperature gradients.

The results validated the adopted solution. As well as helping to obtain high heat transfer coefficients, the combination of key parameters of each of the thermal extraction technologies helps to improve the uniformity of temperature of the heat sink, or even to achieve specific temperature distributions adapted to the specific needs for other applications.

This study opens the way for the optimization of all the key parameters of the proposed cooling scheme (inlet velocity, channel distribution, the presence of curved surfaces under the impinging jet, channel height, etc ) and towards obtaining a heat sink that provides both a very low thermal resistance and a weak temperature gradient across its base. Experimental; q"eff=32 W/cm2 Energy balance prediction; q"eff= 32 W/cm2

Experimental; q"eff=31 W/cm2 Energy balance prediction; q"eff= 31 W/cm2

Experimental; q"eff=21,8 W/cm2 Energy balance prediction; q"eff= 21,8 W/cm2

Experimental; q"eff=13,2 W/cm2 Energy balance prediction; q"eff= 13,2 W/cm2

  position of the bottom of the heat sink (m) z vertical position (m) the temperature of the bottom of the heat sink (K)

Fig. 1 .

 1 Fig. 1. Hybrid jet impingement /micro-channel cooling system by Sung andMudawar[START_REF] Sung | Experimental and numerical investigation of singlephase heat transfer using a hybrid jet-impingement/micro-channel cooling scheme[END_REF]

Fig. 3 .

 3 Fig. 3. Schemes and photography of the experimental part. a) Front view, b) Upper View, c) Side view, d) photography

Fig. 7 .

 7 Fig. 7. Temperature maps of the experimental heat sink. Every point contents uncertainty bars according to the accuracy of the sensor.

Fig. 8 .

 8 Fig. 8. Temperature uniformity at the bottom of the heat sink as a function of the difference in temperature between the points of fluid inlet and outlet. Points include uncertainty bars due to thermocouples.

Fig. 10 .

 10 Fig. 10. Pressure losses from the cooling system. Error bars due to lack of accuracy of the pressure sensor are included.

Fig. 1 .Fig. 2 .Fig. 3 .Fig. 4 .Fig. 5 .Fig. 6 .Fig. 7 .

 1234567 Figure captionsFig.1. Hybrid jet impingement /micro-channel cooling system by Sung and Mudawar[START_REF] Sung | Experimental and numerical investigation of singlephase heat transfer using a hybrid jet-impingement/micro-channel cooling scheme[END_REF]

Fig. 8 .

 8 Fig. 8. Temperature uniformity at the bottom of the heat sink as a function of the difference in temperature between the points of fluid inlet and outlet. Points include uncertainty bars due to thermocouples.

Fig. 9 .

 9 Fig. 9. Nusselt number distribution on the heat sink bottom wall. Error propagation is included in the calculation of the Nusselt number (1.36%).

Fig. 10 .

 10 Fig. 10. Pressure losses from the cooling system. Error bars due to lack of accuracy of the pressure sensor are included.

Fig. 1 .Fig. 2 .Fig. 3 .Fig. 4 .Fig. 5 .

 12345 Fig. 1. Hybrid jet impingement /micro-channel cooling system by Sung andMudawar[START_REF] Sung | Experimental and numerical investigation of singlephase heat transfer using a hybrid jet-impingement/micro-channel cooling scheme[END_REF]

Fig. 6 .Fig. 7 .Fig. 8 .Fig. 9 .

 6789 Fig. 6. Comparison of experimental data and energy balance predictionsfor water temperature rise from heat sink inlet to outlet

Fig. 10 .

 10 Fig. 10. Pressure losses from the cooling system. Error bars due to lack of accuracy of the pressure sensor are included.

Table 1

 1

	Tests conditions
				TEST 1	TEST 2	TEST 3	TEST 4	TEST 5	TEST 6	TEST 7
	T in	ºC	12.35	12.35	12.59	12.73	12.5	12.7	29.2
	Q	m	3 /s	2.07E-05	2.07E-05	1.32E-05	8.42E-06	1.55E-05	1.55E-05	1.60E-05
	q" eff	W/cm	2	21.8	32	31	31	13.2	31	32

Table 2

 2 Temperature uniformity at the bottom of the heat sink

			Symbol	Unit TEST 1 TEST 2 TEST 3 TEST 4 TEST 5 TEST 6 TEST 7
			∆T in-out	ºC	3.7	6.0	8.0	12.5	3.6	7.9	6.8
			T z0,m	ºC	27.1	35.4	39.4	47.3	24.4	41.0	48.7
	Heat sink	bottom	T z0,min T z0,max ∆T min-max σ T	ºC ºC ºC ºC	25.9 28.5 2.6 1.0	33.6 37.4 3.8 1.5	36.5 42.6 6.1 2.4	43.8 49.8 6.0 2.3	22.8 25.7 2.9 1.2	37.4 43.1 5.7 2.3	46.4 50.7 4.3 1.9

Table 3

 3 Average thermal resistance coefficient and pressure losses

						TEST 1	TEST 2	TEST 3	TEST 4	TEST 5	TEST 6	TEST 7
	Q	m	3 /s		2.07E-05 2.07E-05 1.32E-05 8.42E-06 1.55E-05 1.55E-05 1.60E-05
	∆P in-out		Pa		702	620	318	223	435	410	470
	T W		ºC		29.9	39.5	43.1	50.5	25.9	44.1	52.2
	T in		ºC		12.35	12.35	12.59	12.73	12.5	12.7	29.2
	q" eff	W/cm	2	21.8	32	31	31	13.2	31	32
	R tm	m	2 .K/W	8.1E-05	8.5E-05	9.8E-05	1.2E-05	1.0E-05	1.0E-05	7.2E-05