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Experimental investigation of a direct driven radial compressor for domestic heat pumps
international journal of refrigeration 32 (2009) 1918–1928
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Experimental investigation of a direct driven radial compressor for domestic heat pumps J. Schiffmanna,*, D. Favratb a
Fischer Engineering Solutions AG, Birkenweg 3, CH-3360 Herzogenbuchsee, Switzerland Ecole Polytechnique Fe´de´rale de Lausanne, EPFL STI IGM LENI, Station 9, CH-1015 Lausanne, Switzerland
b
article info
abstract
Article history:
The presence of oil in domestic heat pumps is an obstacle toward higher efficiency,
Received 13 April 2009
particularly for enhanced surface evaporators and for advanced concepts based on two-
Received in revised form
stage cycles. Very compact direct driven radial compressors supported on oil-free bearings
1 July 2009
represent a promising alternative. This paper presents the derivation of the specifications,
Accepted 16 July 2009
the choice for an appropriate refrigerant fluid and the design of a proof of concept proto-
Published online 25 July 2009
type with the various tradeoffs between the impeller characteristics to follow the seasonal heat demand, the bearing and rotordynamics for a stable operation. Heat pump simulation
Keywords:
results, the design of the impeller as well as the layout of the experimental facility and first
Heat pump
experimental results are presented. An impeller with a tip diameter of 20 mm has been
Domestic application
tested at rotational speeds of up to 210 krpm reaching pressure ratios in excess of 3.3 and
Centrifugal compressor
efficiencies above 78%. The refrigerant chosen for this first experimental approach is
R134a
HFC 134a. ª 2009 Elsevier Ltd and IIR. All rights reserved.
Design Compressor Simulation Performance
Etude expe´rimentale sur un compresseur radial entraıˆne´ directement pour les pompes a` chaleur re´sidentielles Mots cle´s : Pompe a` chaleur ; Application domestique ; Compresseur centrifuge ; R134a ; Conception ; Compresseur ; Simulation ; Performance
1.
Introduction
According to the IEA (2007) the worldwide energy fraction consumed by the residential sector accounts for 35% of the
total energy consumption, out of which 75% are used for space and tap water heating, showing that domestic heating represents a considerable share in the overall energy consumption. This share is considerably larger for colder climatic zones.
* Corresponding author. Tel.: þ41 (0) 62 956 88 50; fax: þ41 (0) 62 956 22 00. E-mail address: [email protected] (J. Schiffmann). 0140-7007/$ – see front matter ª 2009 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2009.07.006
international journal of refrigeration 32 (2009) 1918–1928
Nomenclature Roman symbols Mechanical Losses [W] E_ W Electrical Power [W] E_ Grid _ m Mass Flow [kg s1] Impeller Tip Width [m] b4 Impeller Diameter [m] d4 Impeller Tip Clearance [m] etip h Enthalpy [J kg1] Rotational Speed [rpm] nrot P Pressure [Pa] T Pressure [ C] Greek symbols Electrical Efficiency hel Internal Isentropic Compressor Efficiency his–k Mechanical Efficiency hmec P Pressure Ratio [–]
Domestic heating requires relatively low temperature levels (30 C for floor space heating and 60 C for tap water), therefore renewable energy sources (environmental heat) can offer an interesting alternative to fossil fuels. For colder climatic zones heat pumps are identified as a key technology for substantially reducing the primary energy consumption and the emissions of greenhouse gases. Heating and cooling will therefore more and more rely on heat pumping in order to enable a more rational use of energy. The recent introduction of an exergy efficiency indicator in a local law on energy and the data proposed to compare heating technologies clearly illustrate the significant advantage of heat pumps in this respect (Favrat et al., 2008). Burer et al. (2003) also show that the future combination of efficient trigeneration techniques with advanced hybrid fuel cell-gas turbine systems and heat pumps could further improve the prospects for a more rational use of energy. Fig. 1 represents the evolution of the coefficient of performance of domestic heat pumps with different cold sources measured at the Swiss heat pump certification center between 1992 and 2008 (Eschmann, 2008). Currently the evolution of the COP is in a stagnating phase. The last surge in performance has been experienced between 1992 and 1995 when the more efficient scroll compressor was introduced. After more than ten years of stagnation a new surge would be more than welcome! A detailed exergetic analysis of high temperature lift heat pumps by Zehnder (2004) has shown that approximately 50% of the losses occur during the compression stage. Another 20% are generated in the heat exchangers due to the temperature pinch. The remainder of the losses can be associated to the expansion process. One of the key issues for decreasing the losses in the heat pump cycle is therefore to reduce the exhaust temperature of the compressor, i.e. by increasing the compressor efficiency. Another way would be to split the compression process into two or more stages associated with intercooling. This latter proposition known for large scale district heat pumps presents the additional advantage of making better use of the exergy of the liquid refrigerant
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Subscripts air Air in Inlet St1 Stage 1 St2 Stage 2 water Water Acronyms AD Aerodynamic AMB Active Magnetic Bearing AS Aerostatic COP Coefficient of Performance FFB Fluid Film Bearing GWP Global Warming Potential HD Hydrodynamic HS Hydrostatic ODP Ozone Depletion Potential OP Operational Point REB Rolling-Element Bearing
at the condenser exhaust, leading to a further reduction of the expansion losses. Favrat et al. (1997) have demonstrated the advantage of domestic multistage heat pumps experimentally. The main challenge when developing multi-stage cycle domestic heat pumps with the available state of the art compressors is the migrating oil which is transported through the cycle by the refrigerant. Oil is required by volumetric compressors for lubricating the sliding surfaces on one hand and for sealing purposes on the other. The investigations by Zehnder (2004) on the oil migration have shown an unstable situation with a net oil migration flow from the second to the first stage compressor, leaving the second one dry and doomed to failure. A self stabilizing configuration could not be identified. The only way of making sure both compressors are
Fig. 1 – The evolution of the COP of heat pumps with different cold sources (brine & air) measured at the Swiss heat pump certification center from 1992 to 2008 (Source: Eschmann, 2008).
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international journal of refrigeration 32 (2009) 1918–1928
well lubricated is an additional oil management system that measures the oil level in the two compressors and transfers it from one to another if necessary. The oil management between the compressors, however, is not the only issue related to the presence of oil in the hermetic refrigerant loop. Youbi-Idrissi and Bonjour (2008) present an overview of the impact of compressor lubrication oil on the different components of refrigeration cycles. It is usually admitted that the evaporator is the most sensitive component to the presence of oil. The thermodynamic properties of oil and refrigerants are considerably different in terms of evaporation temperature and viscosity. This leads to an accumulation of oil at the end of the evaporator, generally leading to a decrease of the heat transfer coefficient (Zuercher et al., 1998a,b; Spindler and Hahne, 2009) and an increase of pressure drop in tubular evaporators (Bandarra-Filho et al., 2009). Youbi-Idrissi et al. (2003, 2004) have also theoretically studied and experimentally investigated the influence of the presence of oil in the evaporator and came to the conclusion that the difference may reach 10%. It follows that a promising way to go for increasing the domestic heat pump performance is to implement multi-stage cycles and to combine them with efficient and oil-free compressor principles. Dynamic machines such as axial or radial compressors do not require oil as their operating principle does not involve surfaces sliding against each other, except for the bearings. Through experimental investigations Balje (1981) proposed to reduce the prediction of rotational speeds, sizes and the expected performance to a set of non-dimensional variables that depend on the required specifications. His maps show that the relation between the pressure ratios and mass flows required by heat pump applications, make radial compressors more suitable than axial ones and that they are able to reach isentropic efficiencies in excess of 85%. It follows that the combination of radial compressors and oil-free bearing technology represents a high potential for a new performance surge in the segment of domestic heat pumps: Such a system allows the elimination of oil in the hermetic cycle enabling the simple implementation of multi-stage cycles with improved performance expectations compared to single stage cycles (Favrat et al., 1997; Zehnder, 2004). As discussed previously an oil-free system also allows to reduce the losses generated by the presence of oil in the evaporator (Zuercher et al., 1998b; YoubiIdrissi et al., 2004) and extends the possibility to use other types of evaporators.
tested by Zehnder et al. (1998). The most promising cycles have been identified as the following: 1. Single stage cycle with an additional separate cycle for upgrading some of the subcooling heat of the condensed liquid to further heat the tap water. 2. Two superposed separate single stage heat pump cycles. A condenser-evaporator couples the two cycles by transferring heat from the bottoming cycle to the topping cycle. 3. Cycle with a two stage expansion, an economizer flash tank acting as a phase separator or an internal heat exchanger at intermediate pressure and a single stage compressor with intermediate vapor injection (Beeton and Pham, 2003; Zehnder and Favrat, 2002). 4. The same as 3 but replacing the single stage compression with injection port by a two-stage compressor.
2.
Heat pump design
Schiffmann et al. (2005) have analyzed the different concepts in view of designing a domestic, high temperature lift air–water heat pump. They have shown that the COP of solution 2 is the worst due to the internal heat transfer losses in the evaporator-condenser linking the two superposed cycles. Solution 1 and 4 represent similar COPs, however, solution 1 is not well adapted to today’s compressor technology as high pressure ratios occur on the primary cycle, especially if high temperature lifts are required. Solution 3, which could be designated as the two-stage cycle of the poor, has a rather moderate COP, but allows to reach high temperature lifts and higher heat rates as has been shown during experimental investigation by Zehnder and Favrat (2005). Solution 3 and 4 with an internal heat exchanger but no economizer-separator are of particular interest for non-azeotropic refrigerant; the internal heat exchanger instead of the economizer avoids a distillation of the refrigerant-blend. However, injection of a potentially two phase refrigerant into the compressor has to be addressed and a separator might have to be considered. Solution 3 is not suitable with a radial compressor as intermediate injection into the stage is difficult to implement without generating excessive losses. Additionally the high pressure ratios (generally above 10) required for high temperature lift heat pumps are not suitable for single stage radial compressor stages. It follows that solution 4 in combination with an economizer-separator is chosen for the future two-stage tests as it is an elegant solution in terms of components and control and promising in terms of COP (Favrat et al., 1997). Furthermore it allows to partly defrost the evaporator by inversing the cycle and using the economizer as an internal energy source. The two-stage cycle is schematically represented in Fig. 2.
2.1.
Twin-stage heat pump cycles
2.2.
Theoretical and experimental work performed by Zehnder (2004) and Favrat et al. (1997) have shown that multi-stage cycles present higher performance than single stage cycles, especially when high temperature lifts are required. These latter are required for tap water heating but also for retrofit applications, where existing gas or oil burners are replaced with heat pumps. In such cases the heat distribution system is often a high temperature hydraulic radiator system. Several twin stage configurations have been presented and some
Refrigerant choice
The choice of the refrigerant influences both, the required pressure ratio and the mass flows and has a considerable effect on the COP. The criteria, however, are not only of thermodynamic but also of ecological nature. Calm and Hourahan (2007) have presented an extensive classification of the different refrigerants according to their ecological impact, the Ozone Depletion Potential (ODP), and the Global Warming Potential (GWP). Today’s refrigerants belong to the third generation with the emphasis on low ODPs, toxicity and
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international journal of refrigeration 32 (2009) 1918–1928
Water In
Water Out
Throttling 2
Condenser
Compression Stage 2
presents a risk of atmospheric air contaminating the refrigerant loop. Flammability is another, non-negligible issue, especially on prototype installations, where small leakages are difficult to avoid. R134a is therefore the chosen refrigerant for this particular application, although its GWP is rather high (1430). Several refrigerant producers are developing alternatives to the R134a with very similar pressure levels and considerably lower GWP (example: R1234yf).
Economizer- Separator Throttling 1
2.3. Compression Stage 1
Air IN
Air OUT
Evaporator
Fig. 2 – A schematic view of the retained two-stage heat pump cycle with an open economizer-separator.
flammability. The growing awareness of climate change further addresses the GWP, opening the opportunity for natural or very low GWP refrigerants. The choice of the refrigerant for this particular project has been addressed by Schiffmann et al. (2005). A summary of the effect on the major variables of a radial turbocompressor driving a twin-stage heat pump resulting from the choice of the refrigerant is given in Table 1. The prediction of the rotational speeds and of the tip diameter of the two impellers has been performed using the design guide maps given by Balje (1981) for reaching the maximum efficiency. In terms of the COP, the worst choice is R407C which is a non-azeotropic refrigerant, additionally presenting the risk of distillation in the intermediate economizer-separator. Ammonia (R717) requires very high rotational speeds leading to very small impellers. A major advantage of propane would be the low overall pressure ratio at the expense of high rotational speeds and of flammability. Both butane (R600) and iso-butane (R600a) present high COPs and additionally operate at low pressures resulting in lower axial forces. The sub-atmospheric inlet pressure, however,
Table 1 – The effect of the refrigerant choice on the major variables related to the compressor driving the twinstage heat pump. Refr.
d4St1 [mm]
d4St2 [mm]
nrot [krpm]
Pin [MPa]
Ptot [–]
COP [–]
R134a R407C R600 R600a R290 R717
23 18 33 28 16 11
20 13 27 24 13 9
200 295 195 220 410 1050
0.14 0.22 0.05 0.08 0.25 0.2
11.8 12.3 13.6 11.6 8.4 13.6
3.01 2.88 3.12 3.07 2.97 3.09
Specifications
The feasibility study by Schiffmann et al. (2002) has shown that the development of radial turbocompressor driving heat pumps bears considerable technological risks. The main unknowns are whether the impeller is able to reach the required pressure ratios at such a small scale and it is not clear how well oil-free bearings lubricated with refrigerant operate at high rotational speeds. In order to check these issues and also to keep the initial prototype as simple as possible it was decided to build and test a single stage compressor serving as proof of concept before stepping further on. The single stage compressor should be able to fit the specifications of the 1st stage of a twin-stage heat pump. The impeller will require a maximum rotational speed of around 220 krpm (see Table 1) for reaching maximum efficiency according to Balje (1981). The electric motor and the power electronics need to be able to drive the rotor up to this speed with a corresponding power of 2.5 kW at the maximum speed. As the heat pump not only operates at one nominal point but rather on a whole range of air and water temperatures the operational range has been divided into 5 typical operational points. Table 2 represents the specifications for the 1st stage of a two stage air–water heat pump that is able to deliver a heat rate of at least 12 kW at a temperature of 60 C at an ambient air temperature of 12 C.
2.4.
Simulation and expected performance
In order to better understand the behavior of a two-stage heat pump with radial turbocompressors it was decided to generate a model of the cycle. Matlab was used for programming and the thermodynamic modeling was linked to a real gas refrigerant database (Lemmon et al., 2002). The major input parameters of the thermodynamic model are the refrigerant, the nominal heat rate, the temperatures of the air inlet and of the heating water outlet, the pinch in the
Table 2 – A summary of the requirements for some typical operational points for the 1st stage of the heat pump compressor processing R134a refrigerant, for an air– water heat pump delivering a heat rate of 12 kW at 60 8C for an external air temperature of L12 8C. OP Tair Twater Pin PSt1 _ St1 m
[ C] [ C] [MPa] [–] [kg/s]
1
2
3
4
5
12 60 0.144 4.2 0.053
7 55 0.177 3.3 0.043
2 50 0.251 2.4 0.024
7 45 0.302 2 0.016
12 40 0.36 1.7 0.005
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international journal of refrigeration 32 (2009) 1918–1928
5
A12W40 A7W45 A2W50 A−7W55 A−12W60
Overall COP [−]
4.5
4
3.5
3
2.5
2 130
140
150
160
170
180
190
200
Rotational Speed [krpm] Fig. 3 – The external COP as a function of the operational point and of the rotational speed.
210
17 16 15 14
Heat Rate [kW]
evaporator and in the condenser as well as the rotational speed of the compressor unit. It is assumed that the two compressor wheels are running on the same shaft and therefore rotating at the same speed. The model then calculates the refrigerant pressures in the heat exchangers and the intermediate pressure in the economizer according to the delivered mass flows of the two impellers. A simple impeller model presented by Schiffmann et al. (2002) allows the estimation of the compressor’s efficiency and whether it tends to operate in surge or in choke. For a given operation point defined by an air inlet temperature and a water exhaust temperature the compressor will be able to run not only on one single rotational speed but over a whole range. The minimum speed is limited by surge phenomena (an aerodynamic phenomenon of instability that might damage the compressor unit due to large oscillating forces). The maximum speed is limited by choke. Higher rotational speeds are possible but do not make any sense as the mass flow would not increase. The optimum COP will generally be located in between these two speed limits. This regulation of the rotational speed allows capacity modulation and therefore a better matching with the heating rate delivered by the heat pump at each operational point. No inlet guide vanes have been considered at this point although this would allow an increase of the operational range. The small size considered makes their implementation complicated and expensive. Figs. 3 and 4 indicate the predicted external COP (defined as the ratio between the delivered heat rate and the total power consumption including the ventilator and the water pump) and the delivered heat rate as a function of the rotational speed. It is interesting to observe the evolution of the heat rate delivered by the heat pump as a function of the external air temperature: it is maximum for low temperatures and minimum for higher air temperatures. This corresponds to an inversion of the delivered heat rate curve compared to conventional air–water heat pumps driven by constant speed rotary volumetric compressors. The radial compressor will therefore allow an operation with a much higher utilization
13 12 11 A12W40 A7W45 A2W50 A−7W55 A−12W60
10 9 8 7 130
140
150
160
170
180
190
200
210
Rotational Speed [krpm]
Fig. 4 – The heat rate as a function of the operational point and of the rotational speed.
coefficient, i.e. the heat pump will run in a much more continuous manner with less stop & starts than conventional systems. The expected COPs are interesting compared to commercially available heat pumps working on the same operation range. As shown by Schiffmann et al. (2002) the annual COP of such a heat pump operating in a climatic profile of Zuerich would be around 3.3. This represents a significant increase compared to commercially available heat pump systems (single stage scroll compressor cycle with intermediate liquid injection).
3.
Compressor unit design
3.1.
Impeller design
The application of a centrifugal compressor for driving a domestic heat pump clearly demands a wide range of pressure ratios and of mass flows. The requirement in terms of efficiency can be stated as the minimization of the energy consumption over a whole heating season. This does not necessarily mean a maximum efficiency on the complete range but rather an efficiency distribution such that the yearly energy consumption of the heat pump system is minimized. Therefore operational points that do not occur often during the heating season do not have an impact as high as frequent ones. A first design of the impeller for the first stage of a domestic two-stage heat pump compressor has been performed by Schiffmann and Favrat (2005) using a commercial design package. As a limitation in terms of size, the minimum impeller diameter was set to 20 mm in order to keep the impeller machinable on a conventional five-axis milling machine. One of the issues related to the design of this particular compressor stage is that it does not operate on one well defined nominal point but rather on the whole operational range with a very wide mass flow range. Variable inlet
international journal of refrigeration 32 (2009) 1918–1928
guide vanes and variable pitch channel diffuser would allow to maximize the efficiency of the stage. Considering the size of the impeller, however, and also the aim of the proof of concept it was decided not to use any variable geometry device for this first prototype. In order to manage the wide mass flow range required by the heat pump it was necessary to use a vaneless diffuser instead of a more efficient fixed geometry channel or even airfoil diffuser. Compared to channel diffusers the vaneless diffuser decreases the efficiency by approximately 2–4 points (Japikse, 1996). Using the 2D and subsequently the 3D tool the impeller geometry was improved. In order to do so particular attention was given to the relative Mach number distribution and to the suppression of secondary flows. The secondary flows are known to move the low momentum fluid of the boundary layers toward regions where the static pressure is low. They are responsible for the jet-wake mechanism resulting in an exit flow non-uniformity known to deteriorate the stage efficiency and reduce the stable operating range of the diffuser (Cumpsty, 2004). Zangeneh et al. (1988a) relate the generation of the secondary flows to the relative Mach number gradients: secondary flow generation occurs whenever there is a gradient of relative Mach numbers in the direction of vorticity in the flow field. Very little secondary flows are therefore generated in the inducer region due to the low boundary layer thickness. The region of high impact is the suction surface in the range of 30–80% of the blade channel length due to the high boundary layer thickness compared to the one on the pressure surface and in the inducer region. In order to avoid secondary flow generation in this region Zangeneh et al. (1988b) propose to reduce the relative Mach number gradients between the hub and the shroud on the suction surface. This mechanism has been applied for the detailed blade geometry design of this particular impeller. Fig. 5 represents the amplitude of the relative Mach number distribution in the streamwise direction at the impeller discharge for the early and for the final design. The surface of the wake is highlighted by the dashed lines. Between the two
1923
designs the surface of the wake has been decreased by a factor 2. At the operation point 3 the predicted stage efficiency has been increased from 0.77 to 0.84 by applying the method described above. Fig. 6 represents the predicted isentropic efficiencies (total-total) and the speed range for the operational points 1, 2, 3 and 5, as a function of the required pressure ratios and mass flows. The prediction is based on 3D CFD calculations. The maximum mass flow at maximum pressure ratio corresponds to a value slightly inferior to the one stated in the specifications; this results from a compromise made to maximize the mass flow at high pressure ratio and to obtain a minimum mass flow corresponding to the specification for the most common operation point (3). The designed impeller allows a continuous operation whenever the air temperature at the inlet of the evaporator is below 2 C (Operation Points 1, 2 and 3). If the air temperature climbs above 2 C then an intermittent compressor operation mode will be necessary, as too much mass flow and therefore too much heating power is delivered. The 3D CFD calculations predict maximum isentropic efficiency in excess of 84%. Mechanical stress calculation as well as the prediction of the blade critical frequencies have been performed using the Finite Element Method provided by a commercial software package (Schiffmann and Favrat, 2005). As the tip speed is low (262 m/s) compared to compressors operating with lighter gases than refrigerants, the mechanical stresses are not a limiting parameter and therefore an easy to machine, conventional aluminum-alloy can be used. The final impeller was machined on a conventional 5 axis milling machine (Fig. 7).
3.2.
Bearing design
The operational environment and the specifications of the machine such as maximum rotational speed, loads, stiffness, damping, or temperature range, just to name a few, have an important impact on the choice of the bearing type. The main bearing technologies that are used in high speed spindle systems can be subdivided into the following categories:
Fig. 5 – The relative Mach number in the streamwise direction at the impeller discharge for the early and for the final design. The surface enclosed by the dashed lines represents the wake for the two designs.
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A−12, Π=4.2 0.85
0.75
0.7
0.7
0.65
0.65
0.6
0
0.025
rpm 5k 20
0.8
195 krpm
0.75
rpm 0k 19
0.85
rpm 0k 21 rpm k 5 21
0.8
ηis−k [−]
A−7, Π=3.4
0.05
0.075
0.1
0.6
178 krpm
0
0.025
0.05
Mass Flow [kg/s]
A2, Π=2.4
A12, Π=1.7
rpm
7k
rpm 5k 18
0.8
0.8 113 krpm
0.75
0.7
0.65
0.65
0
0.025
0.05
rpm 5k 16
0.75
154 krpm
0.7
0.6
0.075
0.1
rpm 9k 13
0.85
16
0.85
ηis−k [−]
0.075
Mass Flow [kg/s]
0.1
Mass Flow [kg/s]
0.6
0
0.025
0.05
0.075
0.1
Mass Flow [kg/s]
Fig. 6 – The isentropic efficiencies and the rotational speed range as a function of the required pressure ratio and mass flow (final geometry).
Rolling-Element Bearings Magnetic Bearings Fluid Bearings In order to enable the choice of the bearing technology to be applied in this particular work an evaluation matrix (Table 3) comparing the different bearing properties has been produced, allowing to identify the most promising bearing technology to be applied in this particular project:
Fig. 7 – The final impeller.
REB: Rolling-element bearings are to be avoided in this project due to their limited life time and due to the need of oil/ grease lubrication, leading to a contamination of the refrigerant in the heat pump loop. Instead of oil or grease the lubricant could also be the refrigerant in liquid phase as proposed by Molyneaux and Zanelli (1996). The design of such a lubrication system, however, would be tedious in the sense that liquid refrigerant has to be continuously provided by the loop, representing a bypass energy. The limited life expectation of the roller bearing operating at high speeds would not get increased either. AMB: Magnetic bearings certainly are a very interesting technology in this field, especially as the compressor unit is a static application and as the operating conditions are stable over a long period of time. The complete bearing system, however is very large compared to other technologies, it requires an auxiliary feedback control and touchdown bearings. The system would become too complex and most likely too expensive. Passive magnetic bearings are unstable and require the integration of a second, stabilizing bearing technology and therefore do not represent a very interesting alternative. Superconductive bearings need an important capital investment in auxiliary cooling units for maintaining temperatures around 70 K. This solution certainly does not represent an economically viable solution for the application in question. FFB: Fluid film bearings are ideal for the high rotational speeds required by the application. Static fluid film bearings
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Table 3 – A summary of the main rotordynamic properties, advantages, drawbacks and required auxiliary systems for the different bearing technologies. C [ Advantageous, H [ Average properties, B [ Not appropriate. Load capacity Contamination Rotordynamics Auxiliaries Wear Losses Temperature Space Cost & availability C H H H B B
B
H C H H
C C
H
B
C B
B
H
B
B
C C H C H
H H H C
require an external source of pressurized gas or liquid, needing a pump or a compressor. These latter represent an additional risk of potential failure, a source of maintenance and a net energy consumption, reducing the total compressor efficiency. Dynamic fluid film bearings require no auxiliary systems and are therefore more reliable. As the heat pump will operate in a continuous manner during the heating season, only very little starts–stops will occur, thus avoiding any damage related to the rubbing at lowest speeds. To be able to design the compressor system as a hermetic unit, the lubrication gas or liquid has to be the same refrigerant as the one of the heat pump loop in order to avoid any contamination, or else complicated sealing systems would be required. It becomes clear that neither Rolling-Element nor Magnetic Bearings correspond to an ideal choice for the system to be designed. REBs due to their limited life expectation and contamination of the heat pump cycle, and AMBs due to their complexity and requirement of auxiliaries. It follows that the solution has to be based on fluid film bearings; ideally gas lubricated rather than liquid ones, as the integration of the latter would require seals and lead to higher losses. The gas lubricated fluid film bearing should be of dynamic type to avoid any energy consuming auxiliaries. In order to be able to accurately predict the bearing properties as a function of the bearing geometry and operating conditions Schiffmann (2008) has implemented an efficient overall model that includes the journal and the axial bearings. The classical lubrication theory is often based on the perfect gas assumption and does not take into account for slip flow. As the bearings are lubricated with vapor phase refrigerant operating very close to the saturation line the perfect gas assumption is far from being valid. Therefore the effects related to rarefaction and to real gas properties have been implemented into the bearing model. Schiffmann and Favrat (2006) have shown that both, the real gas and the rarefaction effects are significant. Furthermore it has been demonstrated by Schiffmann (2008) that the influence of real gas effects decreases with increasing degree of overheat. As a variation of the bearing clearance might affect the bearing properties the bearing model equally takes into account the radial clearance distortion resulting from centrifugal growth and from thermal effects. The overall bearing model has been implemented in Matlab. The function allows to predict the stiffness and the damping coefficients of the journal bearings for radial and for tilting motion and of the axial bearing for axial and for tilting motion as a function of the respective geometry, of the rotational speed and of the excitation frequency by solving the linearized Reynolds equation.
3.3.
C C B B
H C
B
C
C
H H H C C
B
B
H H H H
H H H H
Rotor design
In order to fulfill all the functions required by the heat pump compressor, the rotor has to be composed of the electric motor transmitting the mechanical power to the shaft, of two radial bearings and a two-sided axial bearing restricting the appropriate degrees of freedom and of two radial impellers for the two compressor stages. Numerous combinations are possible. Only an appropriate rotordynamic model of a gas bearing supported rotor will allow to predict the critical speed and the corresponding stability margins. A general model of such a rotor is presented in Fig. 8. The model allows the simulation of an unsymmetrical, rigid rotor, supported by two soft/rigid mounted gas bearings in view of predicting the critical speeds and the corresponding stability margins. The gas bearing properties are represented by a set of radial and tilting springs and dampers. The two stationary bushings are supported by the casing (base) of the machine, modeled through support springs and dampers. In addition to the contributions of the two journals, axial and tilting springs and dampers are introduced on the rotor through the axial bearing. The variables related to the rotor geometry required for predicting the rigid-body modes and the corresponding stability margins are the mass, the polar and the transverse inertia as well as the distances of the bearing mid-planes to the center of gravity of the rotor. The mass and the transverse inertia of the bushings is required as well. To enable the calculation of the critical speeds and the corresponding stability margins, the system’s equations of motion have to be described. This is achieved by summing the forces and couples acting on the rotor resulting from
la
lb
Rotor C.o.G
REB AMB FFB, HS FFB, HD FFB, AS FFB, AD
Bearing Stiffness & Damping Bearing Support Stiffness & Damping Base
Fig. 8 – The rotor model including the bearing stiffness and damping as well as the soft bearing support.
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Test rig
Experimental results
As already explained earlier, the test rig has been designed such that the compressor inlet pressure is controlled through the temperature level of the cooling water, as the cycle cooler is
TT
PT
FT
02
12
21
DT Filter
DT 32
ET
M
41
DT
PT 13
TT
33
Bearing Aeration
06
Main Cycle
31
Motor Cooling
In order to allow the measurement of the performance of the single stage compressor over its complete operational range a test rig has been designed and built. The test rig is composed of two cycles: the main refrigerant cycle through which the major portion of the flow is conveyed, and the bearing aeration cycle which supplies the bearing section with clean and filtered vapor phase refrigerant. The main cycle is composed of an electronically controllable expansion valve, a heat exchanger and a separator that avoids any droplets or debris flowing into the compressor intake. The role of the heat exchanger is to evacuate the energy brought into the cycle by the compressor. The exchanger is partially filled with liquid refrigerant, therefore allowing to control the inlet pressure through the cooling temperature and additionally to stabilize the compressor inlet pressure. The exhaust pressure is controlled over the rotational speed and over the position of the expansion valve. The compressor unit is connected to an additional external circuit to provide for the cooling of the electric motor. To be able to measure the performance of the compressor unit the test rig has been equipped with temperature, pressure and mass flow probes to measure the thermodynamic properties at the inlet and at the exhaust of the compressor. The compressor itself is equipped with two radial and an axial displacement probes to detect the rotational speed and the radial displacements of the rotor of the compressor. Fig. 9 represents the flow chart of the vapor phase test rig showing the principal components and the location of the measurements. Table 4 summarizes the different transmitters installed on the test rig. The measured data is scanned by two synchronized data acquisition boards, a rapid one for the rotor displacements, the pressures, the mass flow as well as for the electric power measurement and a slow one for the temperatures. The data is visualized in real time using a Virtual Instrument (VI) generated with LabView. The VI allows to plot the evolution of the different thermodynamic as well as the bearing related measurements as functions of time and of the rotational speed. It further plots the measured compressor map and allows to control the electronic expansion valve. A FastFourier-Transform module (FFT) has been integrated as well. The tool allows to detect sub- and/or super-synchronous oscillations, that might occur when the stability margin of the rotor is low or if the compressor is operating close to surge. The module allows to detect these oscillations long before they become audible and enables to measure the frequency and the corresponding amplitude.
5.
PT 11
Separator
4.
TT 01
Compressor
displacements of the rotor in relation to the bushings, and to the absolute base. The detailed modeling and design procedure has been presented by Schiffmann (2008).
Bypass
TT
TT
04
03
Cycle Cooler Water In
Water Out
TT 05
Fig. 9 – The flow chart of the vapor phase test rig showing the principal components and the location of the measurements.
partially filled with liquid refrigerant. Both, the exhaust pressure and the mass flow are regulated through the compressor speed and through a regulated electronic expansion valve. In order to determine the compressor characteristic for a given rotational speed from choke to surge at fixed inlet conditions, the electronic expansion valve was progressively and stepwise closed in order to increase throttling downstream of the
Table 4 – The listing of the different measurements and the corresponding probes. Pos 01 02 03 04 05 06 11 12 13 21 31 32 33 41
Measurement of Inlet Temp. Comp. Exh. Temp. Comp. Inlet Temp. Cooler Refr. Exh. Temp. Cooler Refr. Inlet Temp. Cooler Water Motor Temp. Inlet Pressure Exhaust Pressure Bearing Pressure Mass flow Radial Displacement Comp. Radial Displacement Mot. Axial Displacement Electric Power
Unit
[ C] [ C] [ C] [ C] [ C] [ C] [MPa] [MPa] [MPa] [kg s1] [mm] [mm] [mm] [kW]
1927
international journal of refrigeration 32 (2009) 1918–1928
compressor exhaust. The throttling of the valve was increased _ approached zero, which corresponds to the onset until vP=vm of aerodynamic instabilities like surge or stall. The following equations define several conversion efficiencies related to the compressor unit. his–k represents the internal isentropic efficiency of the compressor stage itself, whereas hmec quantifies the windage losses generated by the rotor and the bearings. The electric efficiency is given by the ratio of the mechanical output of the motor to the electric input from the grid. The ratio includes the electric losses generated by the driver as well as by the electric motor. By neglecting the difference of the kinetic energies of the mass flow between the inlet and the exhaust of the compressor, the definition of the different efficiencies is given as follows: hisk ¼
hmec ¼
hel ¼
Dhis Dhmeas
(1)
_ Dhmeas $m
(2)
_ þ E_ W Dhmeas $m
_ þ E_ W Dhmeas $m E_ Grid
(3)
The overall isentropic-electric efficiency of the compressor unit can be defined as follows: hisel ¼ hisk $hmec $hel ¼
_ Dhis $m E_ Grid
(4)
Fig. 10 represents the measured compressor characteristics and the corresponding measured internal isentropic efficiency his–k for speeds ranging from 150 to 210 krpm with inlet conditions corresponding to an external air temperature of 12 C. A pressure ratio in excess of 3.3 could be reached and the peak internal isentropic efficiency was measured to be more than 79%. The calculated mechanical efficiency hmec varies between 92 and 95%, whereas the electric efficiency was measured to be between 70 and 75%, which is rather low.
The poor electric conversion is due to the oversized power electronics driving the electric motor. As the bearings are of dynamic gas lubricated type the stiffness and the load capacity are limited. Therefore rotors supported on such bearings require high balancing levels in order to minimize the radial displacements due to unbalance. A touchdown of the spinning rotor on the bearing bushings would immediately lead to a bearing failure. On this particular prototype the rotational speed was limited to 210 krpm due to too high radial displacements resulting from unstable and shifting unbalance forces, hence the pressure ratio of 4.2 required by the specifications for the OP A-12 could not be reached yet. Although the top pressure ratio of 4.2 could not be reached, the prototype has demonstrated that a small scale, direct driven turbocompressor on oil-free bearings is feasible. The bearings are completely oil free, lubricated by vapor-phase R134a and have shown their ability to carry the rotor up to 210 krpm. The impeller itself has managed to reach very reasonable pressure ratio with internal isentropic efficiencies in excess of 79%. The permanent magnet electric motor driving the rotor of the compressor consumed a power of 1.8 kW at 210 krpm. The achievement of this first prototype is that it has demonstrated that the main technological risk identified during the feasibility study could be overcome and that all the main components were able to fulfill their task. There is, however, room for improvement. As a matter of fact in order to avoid any risk of touchdown between the impeller and its shroud it was decided to leave a comparatively large relative tip clearance (etip/b4) of approximately 20%. It has been reported by Brasz (1988) that a reduction of the relative tip clearance by 10% allows to increase the isentropic efficiency by 3–4 points. Another huge potential for improvement lies in the development of an appropriate motor driver as the one used for these tests performed with conversion efficiencies around 70% as it was considerably oversized for this particular application.
3.4
70 %
72 %
3.2
74 % 74 %
3 72 %
2.8 76 %
2.6 Π [−]
2.4
70 %
76 %
2.2
78 %
68 %
78 %
21
2 1.8
0 20 0 k krpm rpm 0k rpm 0 17 0 k krpm 16 r 0k pm rpm 18
1.6
15
1.4
0k
rpm
19
1.2 1 20
25
30
35
40 45 Mass Flow [g/s]
Fig. 10 – The measured compressor map.
50
55
60
1928
6.
international journal of refrigeration 32 (2009) 1918–1928
Conclusion
Heat pumps play a major role toward a more rational use of energy in the domain of heating and cooling. The performance of today’s heat pumps is impaired by the presence of oil mixed with the refrigerant and its migration through the hermetic loop, particularly for multi-stage units. Multi-stage cycles have the advantage to enable the design of advanced heat pumps that are able to substitute fuel based heating systems, an application segment the present solutions are not able to cover at the highest possible efficiency. Direct driven turbomachinery has been identified as a suitable technology yielding the potential for higher efficiency. In order to proof the technical feasibility of such a system a single stage compressor unit was designed and built. The oil free gas bearing supported radial turbocompressor with a tip diameter of 20 mm could be tested to speeds up to 210 krpm, reaching pressure ratios higher than 3.3 and powers of 1.8 kW. Internal isentropic compressor efficiencies of up to 79% have been measured. The test unit processed the refrigerant R134a. Hence, the technical feasibility of a small scale, oil-free and direct driven turbocompressor for domestic heat pump applications has been demonstrated.
Acknowledgements The authors gratefully acknowledge the Fischer Precise Group for their support given to this work and for the permission to publish this paper.
references
Balje, O.E., 1981. Turbomachines, a Guide to Design, Selection and Theory. John Wiley & Sons. Bandarra-Filho, E.P., Chen, L., Thome, J.R., 2009. Flow boiling characteristics and flow pattern visualization of refrigerant/ lubricant oil mixtures. International Journal of Refrigeration 32, 185–202. Beeton, W.L., Pham, H.M., 2003. Vapor-injection scroll compressor. ASHRAE Journal 45 (4), 22–27. Brasz, J.J., 1988. Investigation into the effect of tip clearance on centrifugal compressor performance. In: ASME-Paper 88-GT-190. Burer, M., Favrat, D., Tanaka, K., Yamada, K., 2003. Multicriteria optimisation of a district heating cogeneration plant integrating a solid oxyde fuel cell-gas turbine combined cycle, heat pumps and chillers. Energy 28 (6), 497–518. Calm, J.M., Hourahan, G.C., 2007. Refrigerant data update. HPAC Engineering, 50–64. Cumpsty, N.A., 2004. Compressor Aerodynamics. Krieger Publishing Company, Malabar, FL, USA. Eschmann, M., 2008. Monitoring von Klein-Waermepumpen mittels Normpruefungen 2008. Tech. Rep. OFEN. Swiss Federal Institute for Energy. Favrat, D., Nidegger, E., Reymond, D., Courtin, G., 1997. Comparison between a single-stage and a two-stage air to water domestic heat pump with one variable speed compressor. In: IIR Conference on Heat Pump Systems, Energy Efficiency and Global Warming, Linz, Austria. Favrat, D., Marechal, F., Epelly, O., 2008. The challenge of introducing an exergy indicator in a local law on energy. Energy 33 (2), 130–136.
IEA, 2007. Renewables for Heating and Cooling. International Energy Agency. Japikse, D., 1996. Centrifugal Compressor Design and Performance. Concepts ETI, Inc. Lemmon, E., McLinden, M., Huber, M., 2002. NIST Reference Fluid Thermodynamic and Transport Properties, vol. 23. Standard Reference Database. Molyneaux, A., Zanelli, R., 1996. Externally pressurized and hybrid bearings lubricated with R134a for oil-free compressors. In: International Compressor Engineering Conference at Purdue, vol. II, pp. 419–424. Schiffmann, J., 2008. Integrated design, optimization and experimental investigation of a direct driven turbocompressor for domestic heat pumps. Ph.D. thesis. Ecole Polytechnique Federale de Lausanne. Schiffmann, J., Favrat, D., 2005. Theoretical design of a high-speed low power radial turbocompressor. In: 6th European Turbomachinery Conference, vol. 2. Lille, France, pp. RC–057–01/71. Schiffmann, J., Favrat, D., 2006. Multi-objective optimization of herringbone grooved gas bearings supporting a high-speed rotor, taking into account rarefied gas and real gas effects. ESDA-95086. In: 8th Biennial ASME Conference on Engineering Systems Design and Analysis July 4–7, 2006, ASME, Torino, Italy. Schiffmann, J., Molyneaux, A., Favrat, D., Mare´chal, F., Zehnder, M., Godat, J., 2002. Compresseur radial pour pompe a` chaleur bie´tage´e, phase 1. Tech. Rep. OFEN. Swiss Federal Office for Energy. Schiffmann, J., Favrat, D., Molyneaux, A., 2005. Theoretical design of a high speed, oil free radial compressor for domestic heat pumps. In: Proceedings of the 8th International Energy Association Heatpump Conference 2005, Las Vegas, USA. Spindler, K., Hahne, E., 2009. The influence of oil on nucleate pool boiling heat transfer. Heat and Mass Transfer 45, 979–990. Youbi-Idrissi, M., Bonjour, J., 2008. The effect of oil in refrigeration: current research issues and critical review of thermodynamic aspects. International Journal of Refrigeration 31, 165–179. Youbi-Idrissi, M., Bonjour, J., Marvillet, C., Meunier, F., 2003. Impact of refrigerant-oil solubility on an evaporator performances working with R407c. International Journal of Refrigeration 26 (3), 284–292. Youbi-Idrissi, M., Bonjour, J., Terrier, J., Marvillet, C., Meunier, F., 2004. Oil presence in an evaporator: experimental validation of a refrigerant/oil mixture enthalpy calculation model. International Journal of Refrigeration 27 (3), 215–224. Zangeneh, M., Dawes, W.N., Hawthorne, W., 1988a. Three dimensional flow in radial-inflow turbines. ASME 88-GT-70. Zangeneh, M., Goto, A., Harada, H., 1998b. On the design criteria for suppression of secondary flows in centrifugal and mixed flow impellers. Journal of Turbomachinery 120, 723–735. Zehnder, M., 2004. Efficient air–water heat pumps for high temperature lift residential heating, including oil migration aspects. Ph.D. thesis, Ecole Polytechnique Federale de Lausanne. Zehnder, M., Favrat, D., 2002. High performance air–water heat pump with extended application range for residential heating. In: 7th IEA Heat Pump Conference, Beijing, China. Zehnder, M., Favrat, D., 2005. Experiences with hermetic scroll compressors with intermediate vapor injection port in heat pumps for high temperature lift heating applications. In: 8th IEA heat pump conference, Las Vegas, USA. Zehnder, M., Favrat, D., Reiner, G., Brugnoli, C., 1998. Waermepumpe mit Hilfskreislauf zur Kondensatunterkuehlung, phase 1. Tech. Rep. OFEN. Swiss Federal Office for Energy. Zuercher, O., Thome, J.R., Favrat, D., 1998a. Intube flow boiling of R407c and R407c/oil mixtures. Part I: microfin tube. HVAC/R Research ASHRAE 4 (4), 347–372. Zuercher, O., Thome, J.R., Favrat, D., 1998b. Intube flow boiling of R407c and R407c/oil mixtures. Part II: plain tube results and predictions. HVAC/R Research ASHRAE 4 (4), 373–400.
available at www.sciencedirect.com
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journal homepage: www.elsevier.com/locate/ijrefrig
Experimental investigation of a direct driven radial compressor for domestic heat pumps J. Schiffmanna,*, D. Favratb a
Fischer Engineering Solutions AG, Birkenweg 3, CH-3360 Herzogenbuchsee, Switzerland Ecole Polytechnique Fe´de´rale de Lausanne, EPFL STI IGM LENI, Station 9, CH-1015 Lausanne, Switzerland
b
article info
abstract
Article history:
The presence of oil in domestic heat pumps is an obstacle toward higher efficiency,
Received 13 April 2009
particularly for enhanced surface evaporators and for advanced concepts based on two-
Received in revised form
stage cycles. Very compact direct driven radial compressors supported on oil-free bearings
1 July 2009
represent a promising alternative. This paper presents the derivation of the specifications,
Accepted 16 July 2009
the choice for an appropriate refrigerant fluid and the design of a proof of concept proto-
Published online 25 July 2009
type with the various tradeoffs between the impeller characteristics to follow the seasonal heat demand, the bearing and rotordynamics for a stable operation. Heat pump simulation
Keywords:
results, the design of the impeller as well as the layout of the experimental facility and first
Heat pump
experimental results are presented. An impeller with a tip diameter of 20 mm has been
Domestic application
tested at rotational speeds of up to 210 krpm reaching pressure ratios in excess of 3.3 and
Centrifugal compressor
efficiencies above 78%. The refrigerant chosen for this first experimental approach is
R134a
HFC 134a. ª 2009 Elsevier Ltd and IIR. All rights reserved.
Design Compressor Simulation Performance
Etude expe´rimentale sur un compresseur radial entraıˆne´ directement pour les pompes a` chaleur re´sidentielles Mots cle´s : Pompe a` chaleur ; Application domestique ; Compresseur centrifuge ; R134a ; Conception ; Compresseur ; Simulation ; Performance
1.
Introduction
According to the IEA (2007) the worldwide energy fraction consumed by the residential sector accounts for 35% of the
total energy consumption, out of which 75% are used for space and tap water heating, showing that domestic heating represents a considerable share in the overall energy consumption. This share is considerably larger for colder climatic zones.
* Corresponding author. Tel.: þ41 (0) 62 956 88 50; fax: þ41 (0) 62 956 22 00. E-mail address: [email protected] (J. Schiffmann). 0140-7007/$ – see front matter ª 2009 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2009.07.006
international journal of refrigeration 32 (2009) 1918–1928
Nomenclature Roman symbols Mechanical Losses [W] E_ W Electrical Power [W] E_ Grid _ m Mass Flow [kg s1] Impeller Tip Width [m] b4 Impeller Diameter [m] d4 Impeller Tip Clearance [m] etip h Enthalpy [J kg1] Rotational Speed [rpm] nrot P Pressure [Pa] T Pressure [ C] Greek symbols Electrical Efficiency hel Internal Isentropic Compressor Efficiency his–k Mechanical Efficiency hmec P Pressure Ratio [–]
Domestic heating requires relatively low temperature levels (30 C for floor space heating and 60 C for tap water), therefore renewable energy sources (environmental heat) can offer an interesting alternative to fossil fuels. For colder climatic zones heat pumps are identified as a key technology for substantially reducing the primary energy consumption and the emissions of greenhouse gases. Heating and cooling will therefore more and more rely on heat pumping in order to enable a more rational use of energy. The recent introduction of an exergy efficiency indicator in a local law on energy and the data proposed to compare heating technologies clearly illustrate the significant advantage of heat pumps in this respect (Favrat et al., 2008). Burer et al. (2003) also show that the future combination of efficient trigeneration techniques with advanced hybrid fuel cell-gas turbine systems and heat pumps could further improve the prospects for a more rational use of energy. Fig. 1 represents the evolution of the coefficient of performance of domestic heat pumps with different cold sources measured at the Swiss heat pump certification center between 1992 and 2008 (Eschmann, 2008). Currently the evolution of the COP is in a stagnating phase. The last surge in performance has been experienced between 1992 and 1995 when the more efficient scroll compressor was introduced. After more than ten years of stagnation a new surge would be more than welcome! A detailed exergetic analysis of high temperature lift heat pumps by Zehnder (2004) has shown that approximately 50% of the losses occur during the compression stage. Another 20% are generated in the heat exchangers due to the temperature pinch. The remainder of the losses can be associated to the expansion process. One of the key issues for decreasing the losses in the heat pump cycle is therefore to reduce the exhaust temperature of the compressor, i.e. by increasing the compressor efficiency. Another way would be to split the compression process into two or more stages associated with intercooling. This latter proposition known for large scale district heat pumps presents the additional advantage of making better use of the exergy of the liquid refrigerant
1919
Subscripts air Air in Inlet St1 Stage 1 St2 Stage 2 water Water Acronyms AD Aerodynamic AMB Active Magnetic Bearing AS Aerostatic COP Coefficient of Performance FFB Fluid Film Bearing GWP Global Warming Potential HD Hydrodynamic HS Hydrostatic ODP Ozone Depletion Potential OP Operational Point REB Rolling-Element Bearing
at the condenser exhaust, leading to a further reduction of the expansion losses. Favrat et al. (1997) have demonstrated the advantage of domestic multistage heat pumps experimentally. The main challenge when developing multi-stage cycle domestic heat pumps with the available state of the art compressors is the migrating oil which is transported through the cycle by the refrigerant. Oil is required by volumetric compressors for lubricating the sliding surfaces on one hand and for sealing purposes on the other. The investigations by Zehnder (2004) on the oil migration have shown an unstable situation with a net oil migration flow from the second to the first stage compressor, leaving the second one dry and doomed to failure. A self stabilizing configuration could not be identified. The only way of making sure both compressors are
Fig. 1 – The evolution of the COP of heat pumps with different cold sources (brine & air) measured at the Swiss heat pump certification center from 1992 to 2008 (Source: Eschmann, 2008).
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international journal of refrigeration 32 (2009) 1918–1928
well lubricated is an additional oil management system that measures the oil level in the two compressors and transfers it from one to another if necessary. The oil management between the compressors, however, is not the only issue related to the presence of oil in the hermetic refrigerant loop. Youbi-Idrissi and Bonjour (2008) present an overview of the impact of compressor lubrication oil on the different components of refrigeration cycles. It is usually admitted that the evaporator is the most sensitive component to the presence of oil. The thermodynamic properties of oil and refrigerants are considerably different in terms of evaporation temperature and viscosity. This leads to an accumulation of oil at the end of the evaporator, generally leading to a decrease of the heat transfer coefficient (Zuercher et al., 1998a,b; Spindler and Hahne, 2009) and an increase of pressure drop in tubular evaporators (Bandarra-Filho et al., 2009). Youbi-Idrissi et al. (2003, 2004) have also theoretically studied and experimentally investigated the influence of the presence of oil in the evaporator and came to the conclusion that the difference may reach 10%. It follows that a promising way to go for increasing the domestic heat pump performance is to implement multi-stage cycles and to combine them with efficient and oil-free compressor principles. Dynamic machines such as axial or radial compressors do not require oil as their operating principle does not involve surfaces sliding against each other, except for the bearings. Through experimental investigations Balje (1981) proposed to reduce the prediction of rotational speeds, sizes and the expected performance to a set of non-dimensional variables that depend on the required specifications. His maps show that the relation between the pressure ratios and mass flows required by heat pump applications, make radial compressors more suitable than axial ones and that they are able to reach isentropic efficiencies in excess of 85%. It follows that the combination of radial compressors and oil-free bearing technology represents a high potential for a new performance surge in the segment of domestic heat pumps: Such a system allows the elimination of oil in the hermetic cycle enabling the simple implementation of multi-stage cycles with improved performance expectations compared to single stage cycles (Favrat et al., 1997; Zehnder, 2004). As discussed previously an oil-free system also allows to reduce the losses generated by the presence of oil in the evaporator (Zuercher et al., 1998b; YoubiIdrissi et al., 2004) and extends the possibility to use other types of evaporators.
tested by Zehnder et al. (1998). The most promising cycles have been identified as the following: 1. Single stage cycle with an additional separate cycle for upgrading some of the subcooling heat of the condensed liquid to further heat the tap water. 2. Two superposed separate single stage heat pump cycles. A condenser-evaporator couples the two cycles by transferring heat from the bottoming cycle to the topping cycle. 3. Cycle with a two stage expansion, an economizer flash tank acting as a phase separator or an internal heat exchanger at intermediate pressure and a single stage compressor with intermediate vapor injection (Beeton and Pham, 2003; Zehnder and Favrat, 2002). 4. The same as 3 but replacing the single stage compression with injection port by a two-stage compressor.
2.
Heat pump design
Schiffmann et al. (2005) have analyzed the different concepts in view of designing a domestic, high temperature lift air–water heat pump. They have shown that the COP of solution 2 is the worst due to the internal heat transfer losses in the evaporator-condenser linking the two superposed cycles. Solution 1 and 4 represent similar COPs, however, solution 1 is not well adapted to today’s compressor technology as high pressure ratios occur on the primary cycle, especially if high temperature lifts are required. Solution 3, which could be designated as the two-stage cycle of the poor, has a rather moderate COP, but allows to reach high temperature lifts and higher heat rates as has been shown during experimental investigation by Zehnder and Favrat (2005). Solution 3 and 4 with an internal heat exchanger but no economizer-separator are of particular interest for non-azeotropic refrigerant; the internal heat exchanger instead of the economizer avoids a distillation of the refrigerant-blend. However, injection of a potentially two phase refrigerant into the compressor has to be addressed and a separator might have to be considered. Solution 3 is not suitable with a radial compressor as intermediate injection into the stage is difficult to implement without generating excessive losses. Additionally the high pressure ratios (generally above 10) required for high temperature lift heat pumps are not suitable for single stage radial compressor stages. It follows that solution 4 in combination with an economizer-separator is chosen for the future two-stage tests as it is an elegant solution in terms of components and control and promising in terms of COP (Favrat et al., 1997). Furthermore it allows to partly defrost the evaporator by inversing the cycle and using the economizer as an internal energy source. The two-stage cycle is schematically represented in Fig. 2.
2.1.
Twin-stage heat pump cycles
2.2.
Theoretical and experimental work performed by Zehnder (2004) and Favrat et al. (1997) have shown that multi-stage cycles present higher performance than single stage cycles, especially when high temperature lifts are required. These latter are required for tap water heating but also for retrofit applications, where existing gas or oil burners are replaced with heat pumps. In such cases the heat distribution system is often a high temperature hydraulic radiator system. Several twin stage configurations have been presented and some
Refrigerant choice
The choice of the refrigerant influences both, the required pressure ratio and the mass flows and has a considerable effect on the COP. The criteria, however, are not only of thermodynamic but also of ecological nature. Calm and Hourahan (2007) have presented an extensive classification of the different refrigerants according to their ecological impact, the Ozone Depletion Potential (ODP), and the Global Warming Potential (GWP). Today’s refrigerants belong to the third generation with the emphasis on low ODPs, toxicity and
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international journal of refrigeration 32 (2009) 1918–1928
Water In
Water Out
Throttling 2
Condenser
Compression Stage 2
presents a risk of atmospheric air contaminating the refrigerant loop. Flammability is another, non-negligible issue, especially on prototype installations, where small leakages are difficult to avoid. R134a is therefore the chosen refrigerant for this particular application, although its GWP is rather high (1430). Several refrigerant producers are developing alternatives to the R134a with very similar pressure levels and considerably lower GWP (example: R1234yf).
Economizer- Separator Throttling 1
2.3. Compression Stage 1
Air IN
Air OUT
Evaporator
Fig. 2 – A schematic view of the retained two-stage heat pump cycle with an open economizer-separator.
flammability. The growing awareness of climate change further addresses the GWP, opening the opportunity for natural or very low GWP refrigerants. The choice of the refrigerant for this particular project has been addressed by Schiffmann et al. (2005). A summary of the effect on the major variables of a radial turbocompressor driving a twin-stage heat pump resulting from the choice of the refrigerant is given in Table 1. The prediction of the rotational speeds and of the tip diameter of the two impellers has been performed using the design guide maps given by Balje (1981) for reaching the maximum efficiency. In terms of the COP, the worst choice is R407C which is a non-azeotropic refrigerant, additionally presenting the risk of distillation in the intermediate economizer-separator. Ammonia (R717) requires very high rotational speeds leading to very small impellers. A major advantage of propane would be the low overall pressure ratio at the expense of high rotational speeds and of flammability. Both butane (R600) and iso-butane (R600a) present high COPs and additionally operate at low pressures resulting in lower axial forces. The sub-atmospheric inlet pressure, however,
Table 1 – The effect of the refrigerant choice on the major variables related to the compressor driving the twinstage heat pump. Refr.
d4St1 [mm]
d4St2 [mm]
nrot [krpm]
Pin [MPa]
Ptot [–]
COP [–]
R134a R407C R600 R600a R290 R717
23 18 33 28 16 11
20 13 27 24 13 9
200 295 195 220 410 1050
0.14 0.22 0.05 0.08 0.25 0.2
11.8 12.3 13.6 11.6 8.4 13.6
3.01 2.88 3.12 3.07 2.97 3.09
Specifications
The feasibility study by Schiffmann et al. (2002) has shown that the development of radial turbocompressor driving heat pumps bears considerable technological risks. The main unknowns are whether the impeller is able to reach the required pressure ratios at such a small scale and it is not clear how well oil-free bearings lubricated with refrigerant operate at high rotational speeds. In order to check these issues and also to keep the initial prototype as simple as possible it was decided to build and test a single stage compressor serving as proof of concept before stepping further on. The single stage compressor should be able to fit the specifications of the 1st stage of a twin-stage heat pump. The impeller will require a maximum rotational speed of around 220 krpm (see Table 1) for reaching maximum efficiency according to Balje (1981). The electric motor and the power electronics need to be able to drive the rotor up to this speed with a corresponding power of 2.5 kW at the maximum speed. As the heat pump not only operates at one nominal point but rather on a whole range of air and water temperatures the operational range has been divided into 5 typical operational points. Table 2 represents the specifications for the 1st stage of a two stage air–water heat pump that is able to deliver a heat rate of at least 12 kW at a temperature of 60 C at an ambient air temperature of 12 C.
2.4.
Simulation and expected performance
In order to better understand the behavior of a two-stage heat pump with radial turbocompressors it was decided to generate a model of the cycle. Matlab was used for programming and the thermodynamic modeling was linked to a real gas refrigerant database (Lemmon et al., 2002). The major input parameters of the thermodynamic model are the refrigerant, the nominal heat rate, the temperatures of the air inlet and of the heating water outlet, the pinch in the
Table 2 – A summary of the requirements for some typical operational points for the 1st stage of the heat pump compressor processing R134a refrigerant, for an air– water heat pump delivering a heat rate of 12 kW at 60 8C for an external air temperature of L12 8C. OP Tair Twater Pin PSt1 _ St1 m
[ C] [ C] [MPa] [–] [kg/s]
1
2
3
4
5
12 60 0.144 4.2 0.053
7 55 0.177 3.3 0.043
2 50 0.251 2.4 0.024
7 45 0.302 2 0.016
12 40 0.36 1.7 0.005
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international journal of refrigeration 32 (2009) 1918–1928
5
A12W40 A7W45 A2W50 A−7W55 A−12W60
Overall COP [−]
4.5
4
3.5
3
2.5
2 130
140
150
160
170
180
190
200
Rotational Speed [krpm] Fig. 3 – The external COP as a function of the operational point and of the rotational speed.
210
17 16 15 14
Heat Rate [kW]
evaporator and in the condenser as well as the rotational speed of the compressor unit. It is assumed that the two compressor wheels are running on the same shaft and therefore rotating at the same speed. The model then calculates the refrigerant pressures in the heat exchangers and the intermediate pressure in the economizer according to the delivered mass flows of the two impellers. A simple impeller model presented by Schiffmann et al. (2002) allows the estimation of the compressor’s efficiency and whether it tends to operate in surge or in choke. For a given operation point defined by an air inlet temperature and a water exhaust temperature the compressor will be able to run not only on one single rotational speed but over a whole range. The minimum speed is limited by surge phenomena (an aerodynamic phenomenon of instability that might damage the compressor unit due to large oscillating forces). The maximum speed is limited by choke. Higher rotational speeds are possible but do not make any sense as the mass flow would not increase. The optimum COP will generally be located in between these two speed limits. This regulation of the rotational speed allows capacity modulation and therefore a better matching with the heating rate delivered by the heat pump at each operational point. No inlet guide vanes have been considered at this point although this would allow an increase of the operational range. The small size considered makes their implementation complicated and expensive. Figs. 3 and 4 indicate the predicted external COP (defined as the ratio between the delivered heat rate and the total power consumption including the ventilator and the water pump) and the delivered heat rate as a function of the rotational speed. It is interesting to observe the evolution of the heat rate delivered by the heat pump as a function of the external air temperature: it is maximum for low temperatures and minimum for higher air temperatures. This corresponds to an inversion of the delivered heat rate curve compared to conventional air–water heat pumps driven by constant speed rotary volumetric compressors. The radial compressor will therefore allow an operation with a much higher utilization
13 12 11 A12W40 A7W45 A2W50 A−7W55 A−12W60
10 9 8 7 130
140
150
160
170
180
190
200
210
Rotational Speed [krpm]
Fig. 4 – The heat rate as a function of the operational point and of the rotational speed.
coefficient, i.e. the heat pump will run in a much more continuous manner with less stop & starts than conventional systems. The expected COPs are interesting compared to commercially available heat pumps working on the same operation range. As shown by Schiffmann et al. (2002) the annual COP of such a heat pump operating in a climatic profile of Zuerich would be around 3.3. This represents a significant increase compared to commercially available heat pump systems (single stage scroll compressor cycle with intermediate liquid injection).
3.
Compressor unit design
3.1.
Impeller design
The application of a centrifugal compressor for driving a domestic heat pump clearly demands a wide range of pressure ratios and of mass flows. The requirement in terms of efficiency can be stated as the minimization of the energy consumption over a whole heating season. This does not necessarily mean a maximum efficiency on the complete range but rather an efficiency distribution such that the yearly energy consumption of the heat pump system is minimized. Therefore operational points that do not occur often during the heating season do not have an impact as high as frequent ones. A first design of the impeller for the first stage of a domestic two-stage heat pump compressor has been performed by Schiffmann and Favrat (2005) using a commercial design package. As a limitation in terms of size, the minimum impeller diameter was set to 20 mm in order to keep the impeller machinable on a conventional five-axis milling machine. One of the issues related to the design of this particular compressor stage is that it does not operate on one well defined nominal point but rather on the whole operational range with a very wide mass flow range. Variable inlet
international journal of refrigeration 32 (2009) 1918–1928
guide vanes and variable pitch channel diffuser would allow to maximize the efficiency of the stage. Considering the size of the impeller, however, and also the aim of the proof of concept it was decided not to use any variable geometry device for this first prototype. In order to manage the wide mass flow range required by the heat pump it was necessary to use a vaneless diffuser instead of a more efficient fixed geometry channel or even airfoil diffuser. Compared to channel diffusers the vaneless diffuser decreases the efficiency by approximately 2–4 points (Japikse, 1996). Using the 2D and subsequently the 3D tool the impeller geometry was improved. In order to do so particular attention was given to the relative Mach number distribution and to the suppression of secondary flows. The secondary flows are known to move the low momentum fluid of the boundary layers toward regions where the static pressure is low. They are responsible for the jet-wake mechanism resulting in an exit flow non-uniformity known to deteriorate the stage efficiency and reduce the stable operating range of the diffuser (Cumpsty, 2004). Zangeneh et al. (1988a) relate the generation of the secondary flows to the relative Mach number gradients: secondary flow generation occurs whenever there is a gradient of relative Mach numbers in the direction of vorticity in the flow field. Very little secondary flows are therefore generated in the inducer region due to the low boundary layer thickness. The region of high impact is the suction surface in the range of 30–80% of the blade channel length due to the high boundary layer thickness compared to the one on the pressure surface and in the inducer region. In order to avoid secondary flow generation in this region Zangeneh et al. (1988b) propose to reduce the relative Mach number gradients between the hub and the shroud on the suction surface. This mechanism has been applied for the detailed blade geometry design of this particular impeller. Fig. 5 represents the amplitude of the relative Mach number distribution in the streamwise direction at the impeller discharge for the early and for the final design. The surface of the wake is highlighted by the dashed lines. Between the two
1923
designs the surface of the wake has been decreased by a factor 2. At the operation point 3 the predicted stage efficiency has been increased from 0.77 to 0.84 by applying the method described above. Fig. 6 represents the predicted isentropic efficiencies (total-total) and the speed range for the operational points 1, 2, 3 and 5, as a function of the required pressure ratios and mass flows. The prediction is based on 3D CFD calculations. The maximum mass flow at maximum pressure ratio corresponds to a value slightly inferior to the one stated in the specifications; this results from a compromise made to maximize the mass flow at high pressure ratio and to obtain a minimum mass flow corresponding to the specification for the most common operation point (3). The designed impeller allows a continuous operation whenever the air temperature at the inlet of the evaporator is below 2 C (Operation Points 1, 2 and 3). If the air temperature climbs above 2 C then an intermittent compressor operation mode will be necessary, as too much mass flow and therefore too much heating power is delivered. The 3D CFD calculations predict maximum isentropic efficiency in excess of 84%. Mechanical stress calculation as well as the prediction of the blade critical frequencies have been performed using the Finite Element Method provided by a commercial software package (Schiffmann and Favrat, 2005). As the tip speed is low (262 m/s) compared to compressors operating with lighter gases than refrigerants, the mechanical stresses are not a limiting parameter and therefore an easy to machine, conventional aluminum-alloy can be used. The final impeller was machined on a conventional 5 axis milling machine (Fig. 7).
3.2.
Bearing design
The operational environment and the specifications of the machine such as maximum rotational speed, loads, stiffness, damping, or temperature range, just to name a few, have an important impact on the choice of the bearing type. The main bearing technologies that are used in high speed spindle systems can be subdivided into the following categories:
Fig. 5 – The relative Mach number in the streamwise direction at the impeller discharge for the early and for the final design. The surface enclosed by the dashed lines represents the wake for the two designs.
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international journal of refrigeration 32 (2009) 1918–1928
A−12, Π=4.2 0.85
0.75
0.7
0.7
0.65
0.65
0.6
0
0.025
rpm 5k 20
0.8
195 krpm
0.75
rpm 0k 19
0.85
rpm 0k 21 rpm k 5 21
0.8
ηis−k [−]
A−7, Π=3.4
0.05
0.075
0.1
0.6
178 krpm
0
0.025
0.05
Mass Flow [kg/s]
A2, Π=2.4
A12, Π=1.7
rpm
7k
rpm 5k 18
0.8
0.8 113 krpm
0.75
0.7
0.65
0.65
0
0.025
0.05
rpm 5k 16
0.75
154 krpm
0.7
0.6
0.075
0.1
rpm 9k 13
0.85
16
0.85
ηis−k [−]
0.075
Mass Flow [kg/s]
0.1
Mass Flow [kg/s]
0.6
0
0.025
0.05
0.075
0.1
Mass Flow [kg/s]
Fig. 6 – The isentropic efficiencies and the rotational speed range as a function of the required pressure ratio and mass flow (final geometry).
Rolling-Element Bearings Magnetic Bearings Fluid Bearings In order to enable the choice of the bearing technology to be applied in this particular work an evaluation matrix (Table 3) comparing the different bearing properties has been produced, allowing to identify the most promising bearing technology to be applied in this particular project:
Fig. 7 – The final impeller.
REB: Rolling-element bearings are to be avoided in this project due to their limited life time and due to the need of oil/ grease lubrication, leading to a contamination of the refrigerant in the heat pump loop. Instead of oil or grease the lubricant could also be the refrigerant in liquid phase as proposed by Molyneaux and Zanelli (1996). The design of such a lubrication system, however, would be tedious in the sense that liquid refrigerant has to be continuously provided by the loop, representing a bypass energy. The limited life expectation of the roller bearing operating at high speeds would not get increased either. AMB: Magnetic bearings certainly are a very interesting technology in this field, especially as the compressor unit is a static application and as the operating conditions are stable over a long period of time. The complete bearing system, however is very large compared to other technologies, it requires an auxiliary feedback control and touchdown bearings. The system would become too complex and most likely too expensive. Passive magnetic bearings are unstable and require the integration of a second, stabilizing bearing technology and therefore do not represent a very interesting alternative. Superconductive bearings need an important capital investment in auxiliary cooling units for maintaining temperatures around 70 K. This solution certainly does not represent an economically viable solution for the application in question. FFB: Fluid film bearings are ideal for the high rotational speeds required by the application. Static fluid film bearings
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international journal of refrigeration 32 (2009) 1918–1928
Table 3 – A summary of the main rotordynamic properties, advantages, drawbacks and required auxiliary systems for the different bearing technologies. C [ Advantageous, H [ Average properties, B [ Not appropriate. Load capacity Contamination Rotordynamics Auxiliaries Wear Losses Temperature Space Cost & availability C H H H B B
B
H C H H
C C
H
B
C B
B
H
B
B
C C H C H
H H H C
require an external source of pressurized gas or liquid, needing a pump or a compressor. These latter represent an additional risk of potential failure, a source of maintenance and a net energy consumption, reducing the total compressor efficiency. Dynamic fluid film bearings require no auxiliary systems and are therefore more reliable. As the heat pump will operate in a continuous manner during the heating season, only very little starts–stops will occur, thus avoiding any damage related to the rubbing at lowest speeds. To be able to design the compressor system as a hermetic unit, the lubrication gas or liquid has to be the same refrigerant as the one of the heat pump loop in order to avoid any contamination, or else complicated sealing systems would be required. It becomes clear that neither Rolling-Element nor Magnetic Bearings correspond to an ideal choice for the system to be designed. REBs due to their limited life expectation and contamination of the heat pump cycle, and AMBs due to their complexity and requirement of auxiliaries. It follows that the solution has to be based on fluid film bearings; ideally gas lubricated rather than liquid ones, as the integration of the latter would require seals and lead to higher losses. The gas lubricated fluid film bearing should be of dynamic type to avoid any energy consuming auxiliaries. In order to be able to accurately predict the bearing properties as a function of the bearing geometry and operating conditions Schiffmann (2008) has implemented an efficient overall model that includes the journal and the axial bearings. The classical lubrication theory is often based on the perfect gas assumption and does not take into account for slip flow. As the bearings are lubricated with vapor phase refrigerant operating very close to the saturation line the perfect gas assumption is far from being valid. Therefore the effects related to rarefaction and to real gas properties have been implemented into the bearing model. Schiffmann and Favrat (2006) have shown that both, the real gas and the rarefaction effects are significant. Furthermore it has been demonstrated by Schiffmann (2008) that the influence of real gas effects decreases with increasing degree of overheat. As a variation of the bearing clearance might affect the bearing properties the bearing model equally takes into account the radial clearance distortion resulting from centrifugal growth and from thermal effects. The overall bearing model has been implemented in Matlab. The function allows to predict the stiffness and the damping coefficients of the journal bearings for radial and for tilting motion and of the axial bearing for axial and for tilting motion as a function of the respective geometry, of the rotational speed and of the excitation frequency by solving the linearized Reynolds equation.
3.3.
C C B B
H C
B
C
C
H H H C C
B
B
H H H H
H H H H
Rotor design
In order to fulfill all the functions required by the heat pump compressor, the rotor has to be composed of the electric motor transmitting the mechanical power to the shaft, of two radial bearings and a two-sided axial bearing restricting the appropriate degrees of freedom and of two radial impellers for the two compressor stages. Numerous combinations are possible. Only an appropriate rotordynamic model of a gas bearing supported rotor will allow to predict the critical speed and the corresponding stability margins. A general model of such a rotor is presented in Fig. 8. The model allows the simulation of an unsymmetrical, rigid rotor, supported by two soft/rigid mounted gas bearings in view of predicting the critical speeds and the corresponding stability margins. The gas bearing properties are represented by a set of radial and tilting springs and dampers. The two stationary bushings are supported by the casing (base) of the machine, modeled through support springs and dampers. In addition to the contributions of the two journals, axial and tilting springs and dampers are introduced on the rotor through the axial bearing. The variables related to the rotor geometry required for predicting the rigid-body modes and the corresponding stability margins are the mass, the polar and the transverse inertia as well as the distances of the bearing mid-planes to the center of gravity of the rotor. The mass and the transverse inertia of the bushings is required as well. To enable the calculation of the critical speeds and the corresponding stability margins, the system’s equations of motion have to be described. This is achieved by summing the forces and couples acting on the rotor resulting from
la
lb
Rotor C.o.G
REB AMB FFB, HS FFB, HD FFB, AS FFB, AD
Bearing Stiffness & Damping Bearing Support Stiffness & Damping Base
Fig. 8 – The rotor model including the bearing stiffness and damping as well as the soft bearing support.
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international journal of refrigeration 32 (2009) 1918–1928
Test rig
Experimental results
As already explained earlier, the test rig has been designed such that the compressor inlet pressure is controlled through the temperature level of the cooling water, as the cycle cooler is
TT
PT
FT
02
12
21
DT Filter
DT 32
ET
M
41
DT
PT 13
TT
33
Bearing Aeration
06
Main Cycle
31
Motor Cooling
In order to allow the measurement of the performance of the single stage compressor over its complete operational range a test rig has been designed and built. The test rig is composed of two cycles: the main refrigerant cycle through which the major portion of the flow is conveyed, and the bearing aeration cycle which supplies the bearing section with clean and filtered vapor phase refrigerant. The main cycle is composed of an electronically controllable expansion valve, a heat exchanger and a separator that avoids any droplets or debris flowing into the compressor intake. The role of the heat exchanger is to evacuate the energy brought into the cycle by the compressor. The exchanger is partially filled with liquid refrigerant, therefore allowing to control the inlet pressure through the cooling temperature and additionally to stabilize the compressor inlet pressure. The exhaust pressure is controlled over the rotational speed and over the position of the expansion valve. The compressor unit is connected to an additional external circuit to provide for the cooling of the electric motor. To be able to measure the performance of the compressor unit the test rig has been equipped with temperature, pressure and mass flow probes to measure the thermodynamic properties at the inlet and at the exhaust of the compressor. The compressor itself is equipped with two radial and an axial displacement probes to detect the rotational speed and the radial displacements of the rotor of the compressor. Fig. 9 represents the flow chart of the vapor phase test rig showing the principal components and the location of the measurements. Table 4 summarizes the different transmitters installed on the test rig. The measured data is scanned by two synchronized data acquisition boards, a rapid one for the rotor displacements, the pressures, the mass flow as well as for the electric power measurement and a slow one for the temperatures. The data is visualized in real time using a Virtual Instrument (VI) generated with LabView. The VI allows to plot the evolution of the different thermodynamic as well as the bearing related measurements as functions of time and of the rotational speed. It further plots the measured compressor map and allows to control the electronic expansion valve. A FastFourier-Transform module (FFT) has been integrated as well. The tool allows to detect sub- and/or super-synchronous oscillations, that might occur when the stability margin of the rotor is low or if the compressor is operating close to surge. The module allows to detect these oscillations long before they become audible and enables to measure the frequency and the corresponding amplitude.
5.
PT 11
Separator
4.
TT 01
Compressor
displacements of the rotor in relation to the bushings, and to the absolute base. The detailed modeling and design procedure has been presented by Schiffmann (2008).
Bypass
TT
TT
04
03
Cycle Cooler Water In
Water Out
TT 05
Fig. 9 – The flow chart of the vapor phase test rig showing the principal components and the location of the measurements.
partially filled with liquid refrigerant. Both, the exhaust pressure and the mass flow are regulated through the compressor speed and through a regulated electronic expansion valve. In order to determine the compressor characteristic for a given rotational speed from choke to surge at fixed inlet conditions, the electronic expansion valve was progressively and stepwise closed in order to increase throttling downstream of the
Table 4 – The listing of the different measurements and the corresponding probes. Pos 01 02 03 04 05 06 11 12 13 21 31 32 33 41
Measurement of Inlet Temp. Comp. Exh. Temp. Comp. Inlet Temp. Cooler Refr. Exh. Temp. Cooler Refr. Inlet Temp. Cooler Water Motor Temp. Inlet Pressure Exhaust Pressure Bearing Pressure Mass flow Radial Displacement Comp. Radial Displacement Mot. Axial Displacement Electric Power
Unit
[ C] [ C] [ C] [ C] [ C] [ C] [MPa] [MPa] [MPa] [kg s1] [mm] [mm] [mm] [kW]
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international journal of refrigeration 32 (2009) 1918–1928
compressor exhaust. The throttling of the valve was increased _ approached zero, which corresponds to the onset until vP=vm of aerodynamic instabilities like surge or stall. The following equations define several conversion efficiencies related to the compressor unit. his–k represents the internal isentropic efficiency of the compressor stage itself, whereas hmec quantifies the windage losses generated by the rotor and the bearings. The electric efficiency is given by the ratio of the mechanical output of the motor to the electric input from the grid. The ratio includes the electric losses generated by the driver as well as by the electric motor. By neglecting the difference of the kinetic energies of the mass flow between the inlet and the exhaust of the compressor, the definition of the different efficiencies is given as follows: hisk ¼
hmec ¼
hel ¼
Dhis Dhmeas
(1)
_ Dhmeas $m
(2)
_ þ E_ W Dhmeas $m
_ þ E_ W Dhmeas $m E_ Grid
(3)
The overall isentropic-electric efficiency of the compressor unit can be defined as follows: hisel ¼ hisk $hmec $hel ¼
_ Dhis $m E_ Grid
(4)
Fig. 10 represents the measured compressor characteristics and the corresponding measured internal isentropic efficiency his–k for speeds ranging from 150 to 210 krpm with inlet conditions corresponding to an external air temperature of 12 C. A pressure ratio in excess of 3.3 could be reached and the peak internal isentropic efficiency was measured to be more than 79%. The calculated mechanical efficiency hmec varies between 92 and 95%, whereas the electric efficiency was measured to be between 70 and 75%, which is rather low.
The poor electric conversion is due to the oversized power electronics driving the electric motor. As the bearings are of dynamic gas lubricated type the stiffness and the load capacity are limited. Therefore rotors supported on such bearings require high balancing levels in order to minimize the radial displacements due to unbalance. A touchdown of the spinning rotor on the bearing bushings would immediately lead to a bearing failure. On this particular prototype the rotational speed was limited to 210 krpm due to too high radial displacements resulting from unstable and shifting unbalance forces, hence the pressure ratio of 4.2 required by the specifications for the OP A-12 could not be reached yet. Although the top pressure ratio of 4.2 could not be reached, the prototype has demonstrated that a small scale, direct driven turbocompressor on oil-free bearings is feasible. The bearings are completely oil free, lubricated by vapor-phase R134a and have shown their ability to carry the rotor up to 210 krpm. The impeller itself has managed to reach very reasonable pressure ratio with internal isentropic efficiencies in excess of 79%. The permanent magnet electric motor driving the rotor of the compressor consumed a power of 1.8 kW at 210 krpm. The achievement of this first prototype is that it has demonstrated that the main technological risk identified during the feasibility study could be overcome and that all the main components were able to fulfill their task. There is, however, room for improvement. As a matter of fact in order to avoid any risk of touchdown between the impeller and its shroud it was decided to leave a comparatively large relative tip clearance (etip/b4) of approximately 20%. It has been reported by Brasz (1988) that a reduction of the relative tip clearance by 10% allows to increase the isentropic efficiency by 3–4 points. Another huge potential for improvement lies in the development of an appropriate motor driver as the one used for these tests performed with conversion efficiencies around 70% as it was considerably oversized for this particular application.
3.4
70 %
72 %
3.2
74 % 74 %
3 72 %
2.8 76 %
2.6 Π [−]
2.4
70 %
76 %
2.2
78 %
68 %
78 %
21
2 1.8
0 20 0 k krpm rpm 0k rpm 0 17 0 k krpm 16 r 0k pm rpm 18
1.6
15
1.4
0k
rpm
19
1.2 1 20
25
30
35
40 45 Mass Flow [g/s]
Fig. 10 – The measured compressor map.
50
55
60
1928
6.
international journal of refrigeration 32 (2009) 1918–1928
Conclusion
Heat pumps play a major role toward a more rational use of energy in the domain of heating and cooling. The performance of today’s heat pumps is impaired by the presence of oil mixed with the refrigerant and its migration through the hermetic loop, particularly for multi-stage units. Multi-stage cycles have the advantage to enable the design of advanced heat pumps that are able to substitute fuel based heating systems, an application segment the present solutions are not able to cover at the highest possible efficiency. Direct driven turbomachinery has been identified as a suitable technology yielding the potential for higher efficiency. In order to proof the technical feasibility of such a system a single stage compressor unit was designed and built. The oil free gas bearing supported radial turbocompressor with a tip diameter of 20 mm could be tested to speeds up to 210 krpm, reaching pressure ratios higher than 3.3 and powers of 1.8 kW. Internal isentropic compressor efficiencies of up to 79% have been measured. The test unit processed the refrigerant R134a. Hence, the technical feasibility of a small scale, oil-free and direct driven turbocompressor for domestic heat pump applications has been demonstrated.
Acknowledgements The authors gratefully acknowledge the Fischer Precise Group for their support given to this work and for the permission to publish this paper.
references
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