制冷外文文献.docx

1、制冷外文文献Keywords:Steady-state simulation、Semi-empirical model、Domestic refrigerators、Experimental validation 1. Introduction A household refrigerator is composed of a thermally insulated cabinet and a vapor-compression refrigeration loop as shown in Fig. 1. These refrigeration systems, on the whole, c

2、onsume a large amount of energy since hundreds of millions are currently in use, and dozens of millions are coming onto the market every year. An understanding of the operational characteristics of a refrigeration system is vital for any energy optimization study, not only to predict its performance

3、, but also to aid the decision making during the design process. The refrigerator performance is usually assessed by one of the following approaches: (i) simplified calculations based on component characteristics; (ii) component analyzes through commercial CFD packages; and (iii) standardized experi

4、ments. Although the first two techniques play important roles in component design,they do not provide enough information on component matching and system behavior, which is only obtained by testing the refrigeratorin a controlled environment chamber. These tests, however, are time consuming and expe

5、nsive. A faster and less costly alternative is the use of computer models to simulate the thermal- and fluid-dynamic behavior of refrigeration systems. Many mathematical models have been proposed in the past for refrigerator modeling. In one of the earliest studies, Davis and Scott 1 developed a mat

6、hematical model to predict the steady-state component behavior over a range of operating conditions, consisting of individual component sub-models that combined first-principles with a number of empirical parameters obtained from the literature. Simplistic models were used for heat exchangers as the

7、 evaporating and condensing pressures were assumed to be known. The compressor model, on the other hand, considered the in-cylinder compression and the pressure drops in the suction and discharge valves. No model has been provided for the expansiondevice. A few years later, the United States Departm

8、ent of Energy (US DOE) sponsored the development of a steady-state simulation model for household refrigerators, which was intended to be adopted as a reference to establish the energy targets for American household manufacturers 2. Based on the US DOE model, several incremental studies were then ca

9、rried out. First, Abramson et al.3 incorporated a sub-model for the capillary tube-suction line heat exchanger, and the model was adapted to simulate a two-doorCombi refrigerator. Later, Reeves et al. 4 improved the overall computational performance using the e-NTU method for heat exchanger modeling

10、, and polynomial fits for the compressor mass flow rate and power consumption. More recently, Klein et al. 5 proposed a first-principles model for simulating the steady-state behavior of a 230-l all-refrigerator, which comprised the following component sub-models: a natural draft wire-and-tube conde

11、nser, a plate-type roll-bond evaporator, a capillary tube-suction line heat exchanger, a hermetic reciprocating compressor, and an insulated cabinet. The model was established based on the mass, momentum and energy conservation laws, heat transfer equations, equations of state of the working media a

12、nd empirical correlations derived from experimental data.The capillary tube-suction line heat exchanger sub-model was derived from a more sophisticated model 6 through a fractioned factorial design technique, whilst the compressor sub-model was based on curve fittings of calorimetric data. In all of

13、 the studies mentioned above, the models depended on reliable component-level performance data, which required purpose-built experimental facilities for testing each component. In the present study, the required empirical information was gathered directly by testing the refrigerator in a controlled

14、temperature and humidity chamber. In order to do so, the refrigeration system was properly and carefully instrumented to minimize any affect on its performance. The conservation laws were employed to establish the governing equations that describe the system behavior. Each component was modeled usin

15、g a lumped approach, based on physical principles and employing empirical parameters (e.g., heat transfer coefficients and friction factors), adjusted to fit the experimental data. The model showed good agreement with experimental data during the validation exercise. 2. Experimental work The tests w

16、ere performed with a 430-l top-mount frost-free refrigerator, assembled with a hermetic reciprocating compressor, a natural draft wire-and-tube condenser, and a tube-fin no-frost evaporator. The sealed system employed HFC-134a as the working fluid (130 g) and synthetic oil as the lubricant (250 ml).

17、 The air temperatures in the freezer and in the fresh-food compartments were controlled by a thermostat and by a thermostatic damper, respectively. The refrigerator was instrumented and installed inside an environment chamber. T-type thermocouple probes were immersed in the refrigerant flow passage

18、and absolute pressure transducers with a measurement uncertainty of 0.1% of the full scale were installed at seven points along the refrigeration loop, as shown in Fig. 1. A Coriolis-type mass flow meter with a measurement uncertainty of 0.03 kg/h was installed at the compressor discharge. The surro

19、unding air temperature was measured by five thermocouples placed around the refrigerator. The freezer and the fresh-food air temperatures were measured, respectively, by three and six thermocouples placed within these compartments. All T-type thermocouples employed in this study have a measurement u

20、ncertainty of 0.3 C. The compressor and fan power consumption were monitored using a digital power analyzer with a measurement uncertainty of 0.1%. A 112-channel system was employed for data acquisition. Tests were performed before and after the instrumentation setup to check for any discrepancies i

21、n the system performance. Additional adjustments were introduced into the system to allow the obtainment of the desired information. A needle metering valve was installed as an auxiliary expansion device upstream of the capillary tube. The original fixed capacity compressor was replaced by a variabl

22、e capacity compressor. The wall heat loop was by-passed and the defrost heaters were turned off. The thermostaticmechanism of the damper was removed and the aperture was kept constantly opened. The compressor and fan power consumption and speed were controlled and measured independently. The compres

23、sor power consumption and speed were measured with a Yokogawa WT230 power analyzer. The fan speed was measured using infrared light. In total, 13 variables were experimentally studied, seven were geometric characteristics of the system and the other six were operational variables. The geometric char

24、acteristics were varied in different combinations which generated eight different system configurations, as shown in Table 1. Each configuration was tested controlling the following six operational variables: (i) ambient temperature; (ii) compressor speed; (iii) refrigerant charge; (iv) auxiliary ex

25、pansion device opening; (v) fan speed; and (vi) internal heating. A total of 168 tests were performed, approximately 20 tests for each configuration. Independent experimental setups were used to measure the capillary tube inner diameter, the internal volumes of the components, and the cabinet overal

26、l thermal conductance. The range of tested conditions is showed in the pressureenthalpy diagram of Fig. 2. It is worth of note that this dataset can also be used for component analysis. Fig. 1. Schematic representation of the vapor-compression loop. 3. Mathematical model For modeling purposes, the r

27、efrigeration system was divided into five component sub-models: (i) compressor, (ii) capillary tube-suction line heat exchanger, (iii) condenser, (iv) evaporator, and (v) refrigerated cabinet. Each of the component sub-models are described below. More detailed information can be found in 7.3.1. Reci

28、procating compressor In most reciprocating compressors, the entering refrigerant passes successively through the compressor shell, the suction muffler and the suction valve to the compression chamber, where it is expelled through the discharge valve to the discharge muffler. The compressor mass flow

29、 rate equation was based on the volumetric efficiency, , as defined by 8 (1) where w and N are the compressor mass flow rate (kg s) and speed (s), respectively, V is the compression chamber volume (m), and v is the specific volume at the compressor inlet (m kg). The compressor discharge enthalpy, h,

30、 was obtained from an overall energy balance (2) where is the overall compression efficiency. The compression power W was calculated as follows: (3) The compressor heat release rate was calculated by an overall thermal conductance, UA, related to the temperature difference between the discharge line

31、, t, and the surrounding air, t. The compressor volumetric and overall efficiency values and the overall compressor thermal conductance were all fitted to the experimental data, yieldingwhere UA is given in (W K), p and pin (bar), tand tin (C), and N in (rpm). As can be seen in Fig. 3, this set of e

32、quations predicts the experimental data for mass flow rate (Fig. 3a) and power consumption (Fig. 3b) within 10% error bands, and the compressor discharge temperature (Fig. 3c) with deviations of 5 C. Fig. 3. Validation of the compressor sub-model: (a) mass flow rate, (b) power consumption and (c) compressor discharge temperature3.2. Heat exchangers: condenser and evaporator The condenser is a natural draft wire-and-tube heat exchanger, in which the air-side temperature is assumed to be uniform. The condenser was div